U.S. patent number 10,760,519 [Application Number 16/411,030] was granted by the patent office on 2020-09-01 for control device of compression-ignition engine.
This patent grant is currently assigned to Mazda Motor Corporation. The grantee listed for this patent is Mazda Motor Corporation. Invention is credited to Tetsuya Chikada, Atsushi Inoue, Yusuke Kawai, Keiji Maruyama, Tomohiro Nishida, Takuya Ohura, Masanari Sueoka, Tatsuhiro Tokunaga.
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United States Patent |
10,760,519 |
Sueoka , et al. |
September 1, 2020 |
Control device of compression-ignition engine
Abstract
A method of implementing control logic of a compression-ignition
engine is provided. A control part of the engine performs a
calculation according to the control logic corresponding to an
engine operating state in response to a measurement of a
measurement part, controls a fuel injection part, a variable valve
operating mechanism, an ignition part and a supercharger so that a
G/F becomes leaner than a stoichiometric air fuel ratio and a A/F
becomes equal to or richer than the stoichiometric air fuel ratio,
while causing the supercharger to boost, and controls the ignition
part so that unburnt mixture gas combusts by self-ignition after
the ignition. The method includes determining a supercharging
pressure P, and determining control logic defining a close timing
IVC of an intake valve. When determining the control logic, the
close timing IVC (deg.aBDC) is determined so that the supercharging
pressure P (kPa) satisfies the following expression:
P.gtoreq.8.0.times.10.sup.-11IVC.sup.6-1.0.times.10.sup.-8IVC.sup.5+3.0.t-
imes.10.sup.-7IVC.sup.4-4.0.times.10.sup.-6IVC.sup.3+0.0068IVC.sup.2-0.320-
9IVC+116.63.
Inventors: |
Sueoka; Masanari (Hiroshima,
JP), Inoue; Atsushi (Aki-gun, JP),
Maruyama; Keiji (Hiroshima, JP), Ohura; Takuya
(Hiroshima, JP), Nishida; Tomohiro (Hiroshima,
JP), Kawai; Yusuke (Hiroshima, JP),
Chikada; Tetsuya (Higashihiroshima, JP), Tokunaga;
Tatsuhiro (Aki-gun, JP) |
Applicant: |
Name |
City |
State |
Country |
Type |
Mazda Motor Corporation |
Aki-gun, Hiroshima |
N/A |
JP |
|
|
Assignee: |
Mazda Motor Corporation
(Aki-gun, Hiroshima, JP)
|
Family
ID: |
66589272 |
Appl.
No.: |
16/411,030 |
Filed: |
May 13, 2019 |
Prior Publication Data
|
|
|
|
Document
Identifier |
Publication Date |
|
US 20190360422 A1 |
Nov 28, 2019 |
|
Foreign Application Priority Data
|
|
|
|
|
May 22, 2018 [JP] |
|
|
2018-098134 |
May 22, 2018 [JP] |
|
|
2018-098136 |
|
Current U.S.
Class: |
1/1 |
Current CPC
Class: |
F02D
35/023 (20130101); F02B 39/12 (20130101); F02D
37/02 (20130101); F02D 41/0007 (20130101); F02D
13/0261 (20130101); F02D 41/006 (20130101); F02M
26/03 (20160201); F02D 41/2412 (20130101); F02D
41/005 (20130101); F02P 5/045 (20130101); F01L
1/22 (20130101); F02B 11/00 (20130101); F02D
41/0047 (20130101); F02P 9/00 (20130101); F02B
3/08 (20130101); F02P 5/1516 (20130101); F02B
33/36 (20130101); F02D 41/3041 (20130101); F02B
39/16 (20130101); F02D 13/0238 (20130101); F02M
26/01 (20160201); F02D 41/2422 (20130101); F02D
13/0219 (20130101); Y02T 10/12 (20130101); F01N
3/021 (20130101); F02D 2041/001 (20130101); F02D
2250/36 (20130101); Y02T 10/40 (20130101); F02D
2200/0414 (20130101); F01N 3/101 (20130101); F02D
2041/0015 (20130101); F02B 31/085 (20130101); F02D
2041/389 (20130101); F02D 2041/3052 (20130101); F02D
2200/021 (20130101); F01N 3/035 (20130101); F02D
35/026 (20130101); F02D 35/028 (20130101) |
Current International
Class: |
F02D
41/30 (20060101); F02D 37/02 (20060101); F01L
1/22 (20060101); F02B 39/12 (20060101); F02D
41/00 (20060101); F02P 9/00 (20060101); F02B
39/16 (20060101) |
References Cited
[Referenced By]
U.S. Patent Documents
Foreign Patent Documents
|
|
|
|
|
|
|
4082292 |
|
Apr 2008 |
|
JP |
|
2018-087565 |
|
Jun 2018 |
|
JP |
|
Other References
European Patent Office, Extended European Search Report Issued in
Application No. 19174810.2, dated Oct. 25, 2019, Germany, 11 pages.
cited by applicant.
|
Primary Examiner: Bogue; Jesse S
Attorney, Agent or Firm: Alleman Hall Creasman & Tuttle
LLP
Claims
What is claimed is:
1. A method of implementing control logic of a compression-ignition
engine, the engine comprising: a fuel injection part configured to
inject fuel to be supplied in a combustion chamber; a variable
valve operating mechanism configured to change a valve timing of an
intake valve; an ignition part configured to ignite a mixture gas
inside the combustion chamber; a supercharger configured to boost
gas introduced into the combustion chamber; a measurement part
configured to measure a parameter related to an operating state of
the engine; and a control part configured to perform a calculation
according to a control logic corresponding to the operating state
of the engine, in response to the measurement of the measurement
part, and output a signal to the fuel injection part, the variable
valve operating mechanism, the ignition part, and the supercharger,
the control part outputting the signal to the fuel injection part,
the variable valve operating mechanism and the supercharger so that
a gas-fuel ratio (G/F) that is a weight ratio of the entire gas of
the mixture gas inside the combustion chamber to the fuel becomes
leaner than a stoichiometric air fuel ratio, and an air-fuel ratio
(A/F) that is a weight ratio of air contained in the mixture gas to
the fuel becomes the stoichiometric air fuel ratio or richer than
the stoichiometric air fuel ratio, while causing the supercharger
to boost, and outputting the signal to the ignition part so that an
unburnt mixture gas combusts by self-ignition after the ignition
part ignites the mixture gas inside the combustion chamber, the
method of implementing the control logic, comprising the steps of:
determining a supercharging pressure P by the supercharger; and
determining the control logic defining a close timing IVC of the
intake valve, wherein, when determining the control logic, the
close timing IVC (deg.aBDC) is determined so that the supercharging
pressure P (kPa) satisfies the following expression:
P.gtoreq.8.0.times.10.sup.-11IVC.sup.6-1.0.times.10.sup.-8IVC.sup.5+3.0.t-
imes.10.sup.-7IVC.sup.4-4.0.times.10.sup.-6IVC.sup.3+0.0068IVC.sup.2-0.320-
9IVC+116.63.
2. The method of claim 1, wherein the supercharging pressure (kPa)
is determined so as to satisfy P.ltoreq.150.
3. The method of claim 1, wherein the control part determines a
target supercharging pressure corresponding to the operating state
of the engine, and wherein the control part determines the close
timing IVC (deg.aBDC) so that, when the target supercharging
pressure is used as the supercharging pressure P (kPa), the
supercharging pressure P (kPa) satisfies the expression.
4. The method of claim 1, wherein the close timing IVC of the
intake valve changes as the operating state of the engine changes,
and wherein the close timing IVC (deg.aBDC) is determined for each
operating state so that the expression is satisfied.
5. The method of claim 1, wherein the engine operates in a
high-load operating state at a given load or higher.
6. The method of claim 5, wherein the engine operates in a maximum
load operating state.
7. The method of claim 1, wherein a geometric compression ratio
.epsilon. of the engine is set so as to satisfy
10.ltoreq..epsilon.<21.
8. The method of claim 1, wherein the engine is provided with an
exhaust gas recirculation (EGR) system configured to introduce
exhaust gas into the combustion chamber, and wherein the control
part outputs the signal to the EGR system and the ignition part so
that a heat amount ratio used as an index related to a ratio of an
amount of heat generated when the mixture gas inside the combustion
chamber combusts by flame propagation to the entire amount of heat
generated when the mixture gas combusts, becomes a target heat
amount ratio defined corresponding to the operating state of the
engine.
9. The method of claim 8, wherein the control part outputs a signal
to the EGR system and the ignition part so that the heat amount
ratio becomes higher when the load of the engine is higher.
10. A control device for a compression-ignition engine, the engine
comprising: a fuel injection part configured to inject fuel to be
supplied in a combustion chamber; a variable valve operating
mechanism configured to change a valve timing of an intake valve;
an ignition part configured to ignite a mixture gas inside the
combustion chamber; a supercharger configured to boost gas
introduced into the combustion chamber; and a measurement part
configured to measure a parameter related to an operating state of
the engine, the control device comprising: a control part
configured to perform a calculation according to a control logic
corresponding to the operating state of the engine, in response to
the measurement of the measurement part, and output a signal to the
fuel injection part, the variable valve operating mechanism, the
ignition part, and the supercharger, the control part outputting
the signal to the fuel injection part, the variable valve operating
mechanism and the supercharger so that a gas-fuel ratio (G/F) that
is a weight ratio of the entire gas of the mixture gas inside the
combustion chamber to the fuel becomes leaner than a stoichiometric
air fuel ratio, and an air-fuel ratio (A/F) that is a weight ratio
of air contained in the mixture gas to the fuel becomes the
stoichiometric air fuel ratio or richer than the stoichiometric air
fuel ratio, while causing the supercharger to boost, and outputting
the signal to the ignition part so that an unburnt mixture gas
combusts by self-ignition after the ignition part ignites the
mixture gas inside the combustion chamber, wherein the control part
determines a supercharging pressure P by the supercharger, and
determines a close timing IVC of the intake valve, and wherein the
control part determines, according to the control logic, the close
timing IVC (deg.aBDC) so that the supercharging pressure P (kPa)
satisfies the following expression:
P.gtoreq.8.0.times.10.sup.-11IVC.sup.6-1.0.times.10.sup.-8IVC.sup.5+3.0.t-
imes.10.sup.-7IVC.sup.4-4.0.times.10.sup.-6IVC.sup.3+0.0068IVC.sup.2-0.320-
9IVC+116.63.
11. The control device of claim 10, wherein the supercharging
pressure (kPa) is determined so as to satisfy P.ltoreq.150.
12. The control device of claim 10, wherein the control part
determines a target supercharging pressure corresponding to the
operating state of the engine, and wherein the control part
determines the close timing IVC (deg.aBDC) so that, when the target
supercharging pressure is used as the supercharging pressure P
(kPa), the supercharging pressure P (kPa) satisfies the
expression.
13. The control device of claim 10, wherein the close timing IVC of
the intake valve changes as the operating state of the engine
changes, and wherein the close timing IVC (deg.aBDC) is determined
for each operating state so that the expression is satisfied.
14. The control device of claim 10, wherein the engine operates in
a high-load operating state at a given load or higher.
15. The control device of claim 14, wherein the engine operates in
a maximum load operating state.
16. The control device of claim 10, wherein the geometric
compression ratio .epsilon. of the engine is set so as to satisfy
10.ltoreq..epsilon.<21.
17. The control device of claim 10, wherein the engine is provided
with an EGR system configured to introduce exhaust gas into the
combustion chamber, and wherein the control part outputs the signal
to the EGR system and the ignition part so that a heat amount ratio
used as an index related to a ratio of an amount of heat generated
when the mixture gas inside the combustion chamber combusts by
flame propagation to the entire amount of heat generated when the
mixture gas combusts, becomes a target heat amount ratio defined
corresponding to the operating state of the engine.
18. The control device of claim 17, wherein the control part
outputs a signal to the EGR system and the ignition part so that
the heat amount ratio becomes higher when the load of the engine is
higher.
Description
TECHNICAL FIELD
The technology disclosed herein relates to a control device of a
compression-ignition engine.
BACKGROUND OF THE DISCLOSURE
It is known that combustion by compressed self-ignition, in which a
mixture gas combusts instantly without flame propagation, maximizes
fuel efficiency since the combustion period is minimal. However,
various problems must be solved for automobile engines with regard
to combustion by compressed self-ignition. For example, since the
operating states and the environmental conditions vary greatly in
automotive applications, stabilizing compressed self-ignition is a
major problem. The combustion by compressed self-ignition has not
been put to practical use for the automobile engine yet. In order
to solve the problem, for example, JP4,082,292B2 proposes that an
ignition plug ignites the mixture gas, when compressed
self-ignition hardly occurs because of a low combustion-chamber
temperature. By igniting the mixture gas immediately before the
compression top dead center, the pressure around the ignition plug
increases to facilitate the compressed self-ignition.
Unlike the technology disclosed in JP4,082,292B2 in which the
compressed self-ignition is assisted by the ignition of the
ignition plug, the present applicant rather proposes SPCCI (SPark
Controlled Compression Ignition) combustion which is a combination
of SI (Spark Ignition) combustion and CI (Compression Ignition)
combustion. The SI combustion is combustion accompanied by the
flame propagation initiated by forcibly igniting the mixture gas
inside the combustion chamber. The CI combustion is combustion
initiated by the mixture gas inside the combustion chamber carrying
out the compressed self-ignition. The SPCCI combustion is
combustion in which, when the mixture gas inside the combustion
chamber is forcibly ignited to start the combustion by flame
propagation, the unburnt mixture gas inside the combustion chamber
combusts by the compression-ignition because of a pressure buildup
due to the heat generation and the flame propagation of the SI
combustion. Since the SPCCI combustion includes the CI combustion,
it is one form of "the combustion by compression-ignition."
The CI combustion takes place, when the in-cylinder temperature
reaches an ignition temperature defined by the composition of the
mixture gas. Fuel efficiency can be maximized, if the in-cylinder
temperature reaches the ignition temperature near a compression top
dead center and the CI combustion takes place. The in-cylinder
temperature increases according to the increase in the in-cylinder
pressure. The in-cylinder pressure in the SPCCI combustion is a
result of two pressure buildups of a pressure buildup by the
compression work of the piston in a compression stroke, and a
pressure buildup caused by the heat generation of the SI
combustion.
Here, if the CI combustion takes place near a compression top dead
center because of a high in-cylinder temperature at a compression
starting timing due to a high ambient temperature, etc., the
in-cylinder pressure excessively increases to create excessive
combustion noise. In this case, combustion noise can be reduced if
the ignition timing is retarded. However, if the ignition timing is
retarded, since the CI combustion takes place when the piston falls
considerably in the expansion stroke, fuel efficiency is decreased.
Since the pressure buildup caused by the heat generation of the SI
combustion can be utilized in the SPCCI combustion, for example, it
is effective to lower the effective compression ratio and to reduce
the pressure buildup by the compression work of the piston in order
to achieve both the reduction of combustion noise and improvement
in fuel efficiency. Thus, combustion noise can be kept suitable,
without decreasing fuel efficiency.
In order to put to practical use the engine which performs the
SPCCI combustion, it is necessary to take into consideration other
control factors relevant to the in-cylinder temperature, other than
effective compression ratio. However, since the SPCCI combustion is
a new combustion system, no one has found other control factors
until now.
Since the SPCCI combustion is compression-ignition combustion,
combustion noise tends to be increased. The present inventors found
that it was necessary to adjust the temperature inside the
combustion chamber at the start timing of the CI combustion to a
suitable temperature, in order to achieve a stable SPCCI
combustion, while reducing combustion noise. If the temperature
inside the combustion chamber is low, the ignitability of the CI
combustion falls. Combustion noise increases as the temperature
inside the combustion chamber goes up.
The temperature inside the combustion chamber mainly depends on the
geometric compression ratio of the engine, and the temperature
and/or amount of gas introduced into the combustion chamber.
Properties related to the gas introduced into the combustion
chamber depend on the supercharging pressure of the supercharger,
and the valve timing of the intake valve, especially when boosting.
The optimal supercharging pressure was unknown so far.
SUMMARY OF THE DISCLOSURE
The present inventors succeeded in finding a relation between the
supercharging pressure and the close timing of the intake valve,
which appropriately causes SPCCI combustion, as a result of
repeated and diligent examinations of SPCCI combustion. The present
inventors came to invent a control device of a compression-ignition
engine based on this knowledge.
Specifically, the technology disclosed herein relates to a control
device of a compression-ignition engine.
The engine includes a fuel injection part configured to inject fuel
to be supplied in a combustion chamber, a variable valve operating
mechanism configured to change a valve timing of an intake valve,
an ignition part configured to ignite a mixture gas inside the
combustion chamber, a supercharger configured to boost gas
introduced into the combustion chamber, a measurement part
configured to measure a parameter related to an operating state of
the engine, and a control part configured to perform a calculation
according to a control logic corresponding to the operating state
of the engine, in response to the measurement of the measurement
part, and output a signal to the fuel injection part, the variable
valve operating mechanism, the ignition part, and the
supercharger.
The control part outputs the signal to the fuel injection part, the
variable valve operating mechanism and the supercharger so that a
gas-fuel ratio (G/F) that is a weight ratio of the entire gas of
the mixture gas inside the combustion chamber to the fuel becomes
leaner than a stoichiometric air fuel ratio, and an air-fuel ratio
(A/F) that is a weight ratio of air contained in the mixture gas to
the fuel becomes the stoichiometric air fuel ratio or richer than
the stoichiometric air fuel ratio, while causing the supercharger
to boost, and outputs the signal to the ignition part so that the
unburnt mixture gas combusts by self-ignition after the ignition
part ignites the mixture gas inside the combustion chamber.
The control part determines a supercharging pressure P by the
supercharger, and determines a close timing IVC of the intake
valve. The control part determines, according to the control logic,
the close timing IVC (deg.aBDC) so that the supercharging pressure
P (kPa) satisfies the following expression.
P.gtoreq.8.0.times.10.sup.-11IVC.sup.6-1.0.times.10.sup.-8IVC.sup.53.0.ti-
mes.10.sup.-7IVC.sup.4-4.0.times.10.sup.-6IVC.sup.3+0.0068IVC.sup.2-0.3209-
IVC+116.63 (A)
The ignition part ignites the mixture gas inside the combustion
chamber in response to the signal from the control part. The
combustion starts by flame propagation and then the unburnt mixture
gas combusts by self-ignition to complete the combustion. That is,
this engine performs the SPCCI (SPark Controlled Compression
Ignition) combustion.
With this engine, the G/F of the mixture gas is made leaner than
the stoichiometric air fuel ratio and the A/F is made to be the
stoichiometric air fuel ratio or richer than the stoichiometric air
fuel ratio. By making the G/F lean, fuel efficiency of the engine
improves, and by making the A/F the stoichiometric air fuel ratio,
emission performance improves by using a catalyst device.
In addition, the boost by the supercharger results in keeping the
G/F of the mixture gas leaner than the stoichiometric air fuel
ratio. This is advantageous in improving the fuel efficiency of the
engine.
When controlling the engine, the control part first determines the
supercharging pressure P by the supercharger. When the
supercharging pressure P is set, the control part determines the
close timing IVC of the intake valve so that the expression (A) is
satisfied. By setting the close timing so as to satisfy the
expression (A), the engine can perform a stable SPCCI combustion
which is a combination of SI (Spark Ignition) combustion and CI
(Compression Ignition) combustion while keeping combustion noise
within the allowable range and the G/F lean environment, even under
various conditions with different situations of the combustion
chamber.
The index for determining the close timing IVC of the intake valve,
which is available when controlling the engine for performing the
SPCCI combustion, has been unknown until now.
The controlling method defines the relationship between the
supercharging pressure P and the close timing IVC of the intake
valve in order to achieve a suitable SPCCI combustion. When
operating the engine, the control part can set the close timing IVC
of the intake valve within the range in which the relationship is
satisfied. Accordingly, the engine for performing the SPCCI
combustion can put to practical use.
The supercharging pressure (kPa) may be determined so as to satisfy
the following expression. P.gtoreq.150 (B)
This is advantageous in using a comparatively small sized
supercharger.
The control part may determine a target supercharging pressure
corresponding to the operating state of the engine. The control
part may determine the close timing IVC (deg.aBDC) so that, when
the target supercharging pressure is used as the supercharging
pressure P (kPa), the supercharging pressure P (kPa) satisfies the
expression.
For example, when the engine load increases and the fuel amount
supplied into the combustion chamber increases, by performing the
boost by the supercharger, the G/F of the mixture gas is leaner
than the stoichiometric air fuel ratio and the A/F is the
stoichiometric air fuel ratio or richer than the stoichiometric air
fuel ratio, which is advantageous in improving fuel efficiency.
The close timing IVC of the intake valve may change as the
operating state of the engine changes, and the close timing IVC
(deg.aBDC) may be determined for each operating state so that one
of the expressions (A) and (B) is satisfied.
Thus, the engine can stably perform the SPCCI combustion in various
operating states.
The engine may operate in a high-load operating state at a given
load or higher.
In general CI combustion, since the pressure fluctuation at the
ignition is relatively large, the combustion noise becomes too
large when the engine load is high, and may exceed the allowable
range. In this regard, in the SPCCI combustion, the SI combustion
is performed at the start of the combustion and the pressure
fluctuation at the ignition in the SI combustion is small. Thus,
combustion noise can be reduced even when the engine load is
high.
The engine may operate in a maximum load operating state. That is
the engine may perform the SPCCI combustion in the maximum load
operating state.
A geometric compression ratio .epsilon. of the engine may be set so
as to satisfy 10.ltoreq..epsilon.<21. In this way, the geometric
compression ratio can be set suitably.
The engine may be provided with an exhaust gas recirculation (EGR)
system configured to introduce exhaust gas into the combustion
chamber. The control part may output the signal to the EGR system
and the ignition part so that a heat amount ratio used as an index
related to a ratio of an amount of heat generated when the mixture
gas inside the combustion chamber combusts by flame propagation to
the entire amount of heat generated when the mixture gas combusts,
becomes a target heat amount ratio defined corresponding to the
operating state of the engine.
The heat amount ratio of the SPCCI combustion is less than 100%.
The heat amount ratio of the combustion mode where the combustion
completes only by flame propagation without the combustion by
compression ignition (i.e., SI combustion) is 100%.
If the heat amount ratio is increased in the SPCCI combustion, the
ratio of the SI combustion increases, which is advantageous in
reducing combustion noise. Whereas, if the heat amount ratio is
lowered in the SPCCI combustion, the ratio of the CI combustion
increases, which is advantageous in improving fuel efficiency. The
heat amount ratio changes by changing the temperature of the
combustion chamber and/or the ignition timing. For example, when
the temperature inside the combustion chamber is high, the CI
combustion starts at an early timing, and the heat amount ratio
becomes low. Further, when the ignition timing is advanced, the SI
combustion starts at an early timing, and the heat amount ratio
becomes high. By the control part outputting signals to the EGR
system and the ignition part so that the heat amount ratio becomes
the target heat amount ratio defined in accordance with the
operating state of the engine, both the reduction of combustion
noise and the improvement of fuel efficiency can be achieved.
The control part may output a signal to the EGR system and the
ignition part so that the heat amount ratio becomes higher when the
load of the engine is higher.
When the engine load increases, the amount of fuel supplied into
the combustion chamber increases and the temperature inside the
combustion chamber becomes high. By increasing the heat amount
ratio of the SPCCI combustion when the engine load is high, the
combustion noise is reduced.
BRIEF DESCRIPTION OF THE DRAWINGS
FIG. 1 is a view illustrating a configuration of an engine.
FIG. 2 is a view illustrating a configuration of a combustion
chamber, where an upper figure corresponds to a plan view of the
combustion chamber, and a lower figure is a cross-sectional view
taken along a line II-II.
FIG. 3 is a plan view illustrating a configuration of the
combustion chamber and an intake system.
FIG. 4 is a block diagram illustrating a configuration of an engine
control device.
FIG. 5 is a graph illustrating a waveform of SPCCI combustion.
FIG. 6 illustrates maps of the engine, where an upper figure is a
map when the engine is warm, a middle figure is a map when the
engine is half warm, and a lower figure is a map when the engine is
cold.
FIG. 7 illustrates the details of the map when the engine is
warm.
FIG. 8 illustrates charts of a fuel injection timing, an ignition
timing, and a combustion waveform in each operating range of the
map of FIG. 7.
FIG. 9 illustrates a layer structure of the engine map.
FIG. 10 is a flowchart illustrating a control process according to
a layer selection of the map.
An upper figure of FIG. 11 is a graph illustrating a relation
between an engine load and an open timing of an intake valve in
Layer 2, and a lower figure thereof is a graph illustrating a
relation between an engine speed and the open timing of the intake
valve in Layer 2.
An upper figure of FIG. 12 is a graph illustrating a relation
between the engine load and the open timing of the intake valve in
Layer 3, a middle figure thereof is a graph illustrating a relation
between the engine load and a close timing of an exhaust valve in
Layer 3, and a lower figure thereof is a graph illustrating a
relation between the engine load, and an overlap period of the
intake valve and the exhaust valve in Layer 3.
FIG. 13 is a flowchart illustrating a process of an operation
control of the engine executed by an ECU.
FIG. 14 illustrates a relation between the engine load and a target
SI ratio.
FIG. 15 is a graph illustrating an occurring range of the SPCCI
combustion versus an exhaust gas recirculation (EGR) rate in Layer
2.
FIG. 16 is one example of a matrix image utilized in order to
determine a relation between a geometric compression ratio and a
close timing of the intake valve where the SPCCI combustion is
possible in Layer 2.
An upper figure of FIG. 17 illustrates a relation between the
geometric compression ratio and the close timing of the intake
valve where the SPCCI combustion is possible in Layer 2 when high
octane fuel is used, and a lower figure thereof illustrates a
relation between the geometric compression ratio and the close
timing of the intake valve where the SPCCI combustion is possible
in Layer 2 when low octane fuel is used.
FIG. 18 is a graph illustrating a range where the SPCCI combustion
is stabilized versus a gas-fuel ratio (G/F) in Layer 3.
FIG. 19 is one example of a matrix image utilized in order to
determine the relation between the geometric compression ratio and
the close timing of the intake valve where the SPCCI combustion is
possible in Layer 3.
An upper figure of FIG. 20 illustrates a relation between the
geometric compression ratio and the close timing of the intake
valve where the SPCCI combustion is possible in Layer 3 when high
octane fuel is used, and a lower figure thereof illustrates a
relation between the geometric compression ratio and the close
timing of the intake valve where the SPCCI combustion is possible
in Layer 3 when low octane fuel is used.
An upper figure of FIG. 21 illustrates a relation between the
geometric compression ratio and the close timing of the intake
valve where the SPCCI combustion is possible in Layer 2 and Layer 3
when high octane fuel is used, and a lower figure thereof
illustrates a relation between the geometric compression ratio and
the close timing of the intake valve where the SPCCI combustion is
possible in Layer 2 and Layer 3 when low octane fuel is used.
FIG. 22 is a flowchart illustrating a procedure of controlling a
compression-ignition engine.
An upper figure of FIG. 23 illustrates a relation between a
supercharging pressure and the close timing of the intake valve
where a G/F lean SPCCI combustion is possible in Layer 2 when the
geometric compression ratio is set relatively high, and a lower
figure thereof illustrates a relation between a supercharging
pressure and the close timing of the intake valve where the G/F
lean SPCCI combustion is possible in Layer 2 when the geometric
compression ratio is set relatively low.
FIG. 24 is a flowchart illustrating a procedure of controlling the
compression-ignition engine.
DETAILED DESCRIPTION OF THE DISCLOSURE
Hereinafter, one embodiment of a method of implementing control
logic of a compression-ignition engine will be described in detail
with reference to the accompanying drawings. The following
description is one example of the engine and the method of
implementing the control logic.
FIG. 1 is a view illustrating a configuration of the
compression-ignition engine. FIG. 2 is a view illustrating a
configuration of a combustion chamber of the engine. FIG. 3 is a
view illustrating a configuration of the combustion chamber and an
intake system. Note that in FIG. 1, an intake side is the left side
in the drawing, and an exhaust side is the right side in the
drawing. In FIGS. 2 and 3, the intake side is the right side in the
drawings, and the exhaust side is the left side in the drawings.
FIG. 4 is a block diagram illustrating a configuration of a control
device of the engine.
An engine 1 is a four-stroke engine which operates by a combustion
chamber 17 repeating an intake stroke, a compression stroke, an
expansion stroke, and an exhaust stroke. The engine 1 is mounted on
an automobile with four wheels. The automobile travels by operating
the engine 1. Fuel of the engine 1 is gasoline in this example. The
fuel may be a liquid fuel containing at least gasoline. The fuel
may be gasoline containing, for example, bioethanol.
(Engine Configuration)
The engine 1 includes a cylinder block 12 and a cylinder head 13
placed thereon. A plurality of cylinders 11 are formed inside the
cylinder block 12. In FIGS. 1 and 2, only one cylinder 11 is
illustrated. The engine 1 is a multi-cylinder engine.
A piston 3 is slidably inserted in each cylinder 11. The pistons 3
are connected with a crankshaft 15 through respective connecting
rods 14. Each piston 3 defines the combustion chamber 17, together
with the cylinder 11 and the cylinder head 13. Note that the term
"combustion chamber" may be used in a broad sense. That is, the
term "combustion chamber" may refer to a space formed by the piston
3, the cylinder 11, and the cylinder head 13, regardless of the
position of the piston 3.
As illustrated in the lower figure of FIG. 2, a lower surface of
the cylinder head 13, i.e., a ceiling surface of the combustion
chamber 17, is comprised of a slope 1311 and a slope 1312. The
slope 1311 is a rising gradient from the intake side toward an
injection axial center X2 of an injector 6 which will be described
later. The slope 1312 is a rising gradient from the exhaust side
toward the injection axial center X2. The ceiling surface of the
combustion chamber 17 is a so-called "pent-roof" shape.
An upper surface of the piston 3 is bulged toward the ceiling
surface of the combustion chamber 17. A cavity 31 is formed in the
upper surface of the piston 3. The cavity 31 is a dent in the upper
surface of the piston 3. The cavity 31 has a shallow pan shape in
this example. The center of the cavity 31 is offset at the exhaust
side with respect to a center axis X1 of the cylinder 11.
A geometric compression ratio .epsilon. of the engine 1 is set so
as to satisfy 1.ltoreq..epsilon.<30, and preferably satisfy
10.ltoreq..epsilon.<21. The engine 1 which will be described
later performs SPCCI (SPark Controlled Compression Ignition)
combustion that is a combination of SI (Spark Ignition) combustion
and CI (Compression Ignition) combustion in a part of operating
ranges. The SPCCI combustion controls the CI combustion using a
heat generation and a pressure buildup by the SI combustion. The
engine 1 is the compression-ignition engine. However, in this
engine 1, temperature of the combustion chamber 17, when the piston
3 is at a compression top dead center (i.e., compression end
temperature), does not need to be increased. In the engine 1, the
geometric compression ratio can be set comparatively low. The low
geometric compression ratio becomes advantageous in reduction of
cooling loss and mechanical loss. For engines using regular
gasoline (a low octane fuel of which octane number is about 91),
the geometric compression ratio of the engine 1 is 14-17, and for
those using high octane gasoline (high octane fuel of which octane
number is about 96), the geometric compression ratio is 15-18.
An intake port 18 is formed in the cylinder head 13 for each
cylinder 11. As illustrated in FIG. 3, each intake port 18 has a
first intake port 181 and a second intake port 182. The intake port
18 communicates with the corresponding combustion chamber 17.
Although the detailed illustration of the intake port 18 is
omitted, it is a so-called "tumble port." That is, the intake port
18 has such a shape that a tumble flow is formed in the combustion
chamber 17.
Each intake valve 21 is disposed in the intake ports 181 and 182.
The intake valve 21 opens and closes a channel between the
combustion chamber 17 and the intake port 181 or 182. The intake
valves 21 are opened and closed at given timings by a valve
operating mechanism. The valve operating mechanism may be a
variable valve operating mechanism which varies the valve timing
and/or valve lift. In this example, as illustrated in FIG. 4, the
variable valve operating mechanism has an intake-side electric S-VT
(Sequential-Valve Timing) 23. The intake-side electric S-VT 23
continuously varies a rotation phase of an intake cam shaft within
a given angle range. The open timing and the close timing of the
intake valve 21 vary continuously. Note that the electric S-VT may
be replaced with a hydraulic S-VT, as the intake valve operating
mechanism.
An exhaust port 19 is also formed in the cylinder head 13 for each
cylinder 11. As illustrated in FIG. 3, each exhaust port 19 also
has a first exhaust port 191 and a second exhaust port 192. The
exhaust port 19 communicates with the corresponding combustion
chamber 17.
Each exhaust valve 22 is disposed in the exhaust ports 191 and 192.
The exhaust valve 22 opens and closes a channel between the
combustion chamber 17 and the exhaust port 191 or 192. The exhaust
valves 22 are opened and closed at a given timing by a valve
operating mechanism. The valve operating mechanism may be a
variable valve operating mechanism which varies the valve timing
and/or valve lift. In this example, as illustrated in FIG. 4, the
variable valve operating mechanism has an exhaust-side electric SVT
24. The exhaust-side electric S-VT 24 continuously varies a
rotation phase of an exhaust cam shaft within a given angle range.
The open timing and the close timing of the exhaust valve 22 change
continuously. Note that the electric S-VT may be replaced with a
hydraulic S-VT, as the exhaust valve operating mechanism.
The intake-side electric S-VT 23 and the exhaust-side electric S-VT
24 adjust the length of an overlap period where both the intake
valve 21 and the exhaust valve 22 are open. If the length of the
overlap period is made longer, the residual gas in the combustion
chamber 17 can be purged. Moreover, by adjusting the length of the
overlap period, internal EGR (Exhaust Gas Recirculation) gas can be
introduced into the combustion chamber 17. An internal EGR system
is comprised of the intake-side electric S-VT 23 and the
exhaust-side electric S-VT 24. Note that the internal EGR system
may not be comprised of the S-VT.
The injector 6 is attached to the cylinder head 13 for each
cylinder 11. Each injector 6 directly injects fuel into the
combustion chamber 17. The injector 6 is one example of a fuel
injection part. The injector 6 is disposed in a valley part of the
pent roof where the slope 1311 and the slope 1312 meet. As
illustrated in FIG. 2, the injection axial center X2 of the
injector 6 is located at the exhaust side of the center axis X1 of
the cylinder 11. The injection axial center X2 of the injector 6 is
parallel to the center axis X1. The injection axial center X2 of
the injector 6 and the center of the cavity 31 are in agreement
with each other. The injector 6 faces the cavity 31. Note that the
injection axial center X2 of the injector 6 may be in agreement
with the center axis X1 of the cylinder 11. In such a
configuration, the injection axial center X2 of the injector 6 and
the center of the cavity 31 may be in agreement with each
other.
Although detailed illustration is omitted, the injector 6 is
comprised of a multi nozzle-port type fuel injection valve having a
plurality of nozzle ports. As illustrated by two-dot chain lines in
FIG. 2, the injector 6 injects fuel so that the fuel spreads
radially from the center of the combustion chamber 17. The injector
6 has ten nozzle ports in this example, and the nozzle port is
disposed so as to be equally spaced in the circumferential
direction.
The injectors 6 are connected to a fuel supply system 61. The fuel
supply system 61 includes a fuel tank 63 configured to store fuel,
and a fuel supply passage 62 which connects the fuel tank 63 to the
injector 6. In the fuel supply passage 62, a fuel pump 65 and a
common rail 64 are provided. The fuel pump 65 pumps fuel to the
common rail 64. The fuel pump 65 is a plunger pump driven by the
crankshaft 15 in this example. The common rail 64 stores fuel
pumped from the fuel pump 65 at a high fuel pressure. When the
injector 6 is opened, the fuel stored in the common rail 64 is
injected into the combustion chamber 17 from the nozzle ports of
the injector 6. The fuel supply system 61 can supply fuel to the
injectors 6 at a high pressure of 30 MPa or higher. The pressure of
fuel supplied to the injector 6 may be changed according to the
operating state of the engine 1. Note that the configuration of the
fuel supply system 61 is not limited to the configuration described
above.
An ignition plug 25 is attached to the cylinder head 13 for each
cylinder 11. The ignition plug 25 forcibly ignites a mixture gas
inside the combustion chamber 17. The ignition plug 25 is disposed
at the intake side of the center axis X1 of the cylinder 11 in this
example. The ignition plug 25 is located between the two intake
ports 181 and 182 of each cylinder. The ignition plug 25 is
attached to the cylinder head 13 so as to incline downwardly toward
the center of the combustion chamber 17. As illustrated in FIG. 2,
the electrode of the ignition plug 25 faces to the inside of the
combustion chamber 17 and is located near the ceiling surface of
the combustion chamber 17. Note that the ignition plug 25 may be
disposed at the exhaust side of the center axis X1 of the cylinder
11. Moreover, the ignition plug 25 may be disposed on the center
axis X1 of the cylinder 11.
An intake passage 40 is connected to one side surface of the engine
1. The intake passage 40 communicates with the intake port 18 of
each cylinder 11. Gas introduced into the combustion chamber 17
flows through the intake passage 40. An air cleaner 41 is disposed
in an upstream end part of the intake passage 40. The air cleaner
41 filters fresh air. A surge tank 42 is disposed near the
downstream end of the intake passage 40. Part of the intake passage
40 downstream of the surge tank 42 constitutes independent passages
branched from the intake passage 40 for each cylinder 11. The
downstream end of each independent passage is connected to the
intake port 18 of each cylinder 11.
A throttle valve 43 is disposed between the air cleaner 41 and the
surge tank 42 in the intake passage 40. The throttle valve 43
adjusts an introducing amount of the fresh air into the combustion
chamber 17 by adjusting an opening of the throttle valve.
A supercharger 44 is also disposed in the intake passage 40,
downstream of the throttle valve 43. The supercharger 44 boosts gas
to be introduced into the combustion chamber 17. In this example,
the supercharger 44 is a mechanical supercharger driven by the
engine 1. The mechanical supercharger 44 may be a root, Lysholm,
vane, or a centrifugal type.
An electromagnetic clutch 45 is provided between the supercharger
44 and the engine 1. The electromagnetic clutch 45 transmits a
driving force from the engine 1 to the supercharger 44 or
disengages the transmission of the driving force between the
supercharger 44 and the engine 1. As will be described later, an
ECU 10 switches the disengagement and engagement of the
electromagnetic clutch 45 to switch the supercharger 44 between ON
and OFF.
An intercooler 46 is disposed downstream of the supercharger 44 in
the intake passage 40. The intercooler 46 cools gas compressed by
the supercharger 44. The intercooler 46 may be of a water cooling
type or an oil cooling type, for example.
A bypass passage 47 is connected to the intake passage 40. The
bypass passage 47 connects an upstream part of the supercharger 44
to a downstream part of the inter cooler 46 in the intake passage
40 so as to bypass the supercharger 44 and the inter cooler 46. An
air bypass valve 48 is disposed in the bypass passage 47. The air
bypass valve 48 adjusts a flow rate of gas flowing in the bypass
passage 47.
The ECU 10 fully opens the air bypass valve 48 when the
supercharger 44 is turned OFF (i.e., when the electromagnetic
clutch 45 is disengaged). The gas flowing through the intake
passage 40 bypasses the supercharger 44 and is introduced into the
combustion chamber 17 of the engine 1. The engine 1 operates in a
non-supercharged state, i.e., a natural aspiration state.
When the supercharger 44 is turned ON, the engine 1 operates in a
supercharged state. The ECU 10 adjusts an opening of the air bypass
valve 48 when the supercharger 44 is turned ON (i.e., when the
electromagnetic clutch 45 is engaged). A portion of the gas which
passed through the supercharger 44 flows back toward upstream of
the supercharger 44 through the bypass passage 47. When the ECU 10
adjusts the opening of the air bypass valve 48, a supercharging
pressure of gas introduced into the combustion chamber 17 changes.
Note that the term "supercharging" as used herein refers to a
situation where the pressure inside the surge tank 42 exceeds an
atmospheric pressure, and "non-supercharging" refers to a situation
where the pressure inside the surge tank 42 becomes below the
atmospheric pressure.
In this example, a supercharging system 49 is comprised of the
supercharger 44, the bypass passage 47, and the air bypass valve
48.
The engine 1 has a swirl generating part which generates a swirl
flow inside the combustion chamber 17. As illustrated in FIG. 3,
the swirl generating part has a swirl control valve 56 attached to
the intake passage 40. Among a primary passage 401 coupled to the
first intake port 181 and a secondary passage 402 coupled to the
second intake port 182, the swirl control valve 56 is disposed in
the secondary passage 402. The swirl control valve 56 is an opening
control valve which is configured to choke a cross section of the
secondary passage 402. When the opening of the swirl control valve
56 is small, since an intake flow rate of air flowing into the
combustion chamber 17 from the first intake port 181 is relatively
large, and an intake flow rate of air flowing into the combustion
chamber 17 from the second intake port 182 is relatively small, the
swirl flow inside the combustion chamber 17 becomes stronger. On
the other hand, when the opening of the swirl control valve 56 is
large, since the intake flow rates of air flowing into the
combustion chamber 17 from the first intake port 181 and the second
intake port 182 become substantially equal, the swirl flow inside
the combustion chamber 17 becomes weaker. When the swirl control
valve 56 is fully opened, the swirl flow will not occur. Note that
the swirl flow circulates counterclockwise in FIG. 3, as
illustrated by white arrows (also see white arrows in FIG. 2).
An exhaust passage 50 is connected to the other side surface of the
engine 1. The exhaust passage 50 communicates with the exhaust port
19 of each cylinder 11. The exhaust passage 50 is a passage through
which exhaust gas discharged from the combustion chambers 17 flows.
Although detailed illustration is omitted, an upstream part of the
exhaust passage 50 constitutes independent passages branched from
the exhaust passage 50 for each cylinder 11. The upper end of the
independent passage is connected to the exhaust port 19 of each
cylinder 11.
An exhaust gas purification system having a plurality of catalytic
converters is disposed in the exhaust passage 50. Although
illustration is omitted, an upstream catalytic converter is
disposed inside an engine room. The upstream catalytic converter
has a three-way catalyst 511 and a GPF (Gasoline Particulate
Filter) 512. The downstream catalytic converter is disposed outside
the engine room. The downstream catalytic converter has a three-way
catalyst 513. Note that the exhaust gas purification system is not
limited to the illustrated configuration. For example, the GPF may
be omitted. Moreover, the catalytic converter is not limited to
those having the three-way catalyst. Further, the order of the
three-way catalyst and the GPF may suitably be changed.
Between the intake passage 40 and the exhaust passage 50, an EGR
passage 52 which constitutes an external EGR system is connected.
The EGR passage 52 is a passage for recirculating part of the
exhaust gas to the intake passage 40. The upstream end of the EGR
passage 52 is connected between the upstream catalytic converter
and the downstream catalytic converter in the exhaust passage 50.
The downstream end of the EGR passage 52 is connected to an
upstream part of the supercharger 44 in the intake passage 40. EGR
gas flowing through the EGR passage 52 flows into the upstream part
of the supercharger 44 in the intake passage 40, without passing
through the air bypass valve 48 of the bypass passage 47.
An EGR cooler 53 of a water cooling type is disposed in the EGR
passage 52. The EGR cooler 53 cools the exhaust gas. An EGR valve
54 is also disposed in the EGR passage 52. The EGR valve 54 adjusts
a flow rate of the exhaust gas flowing through the EGR passage 52.
By adjusting the opening of the EGR valve 54, an amount of the
cooled exhaust gas, i.e., a recirculating amount of external EGR
gas can be adjusted.
In this example, an EGR system 55 is comprised of the external EGR
system and the internal EGR system. The external EGR system can
supply the exhaust gas to the combustion chamber 17 that is lower
in temperature than the internal EGR system.
The control device of the compression-ignition engine includes the
ECU (Engine Control Unit) 10 for operating the engine 1. The ECU 10
is a controller based on a well-known microcomputer, and as
illustrated in FIG. 4, includes a central processing unit (CPU) 101
which executes a computer program, memory 102 which, for example,
is comprised of a RAM (Random Access Memory) and/or a ROM (Read
Only Memory), and stores the program and data, and an input/output
bus 103 which inputs and outputs an electrical signal. The ECU 10
is one example of the control part.
As illustrated in FIGS. 1 and 4, various kinds of sensors SW1-SW17
are connected to the ECU 10. The sensors SW1-SW17 output signals to
the ECU 10. The sensors include the following sensors:
Airflow sensor SW1: Disposed downstream of the air cleaner 41 in
the intake passage 40, and measures a flow rate of fresh air
flowing through the intake passage 40;
First intake-air temperature sensor SW2: Disposed downstream of the
air cleaner 41 in the intake passage 40, and measures the
temperature of fresh air flowing through the intake passage 40;
First pressure sensor SW3: Disposed downstream of the connected
position of the EGR passage 52 in the intake passage 40 and
upstream of the supercharger 44, and measures the pressure of gas
flowing into the supercharger 44;
Second intake-air temperature sensor SW4: Disposed downstream of
the supercharger 44 in the intake passage 40 and upstream of the
connected position of the bypass passage 47, and measures the
temperature of gas flowed out of the supercharger 44;
Second pressure sensor SW5: Attached to the surge tank 42, and
measures the pressure of gas downstream of the supercharger 44;
Pressure indicating sensors SW6: Attached to the cylinder head 13
corresponding to each cylinder 11, and measures the pressure inside
each combustion chamber 17;
Exhaust temperature sensor SW7: Disposed in the exhaust passage 50,
and measures the temperature of the exhaust gas discharged from the
combustion chamber 17;
Linear O.sub.2 sensor SW8: Disposed upstream of the upstream
catalytic converter in the exhaust passage 50, and measures the
oxygen concentration of the exhaust gas;
Lambda O.sub.2 sensor SW9: Disposed downstream of the three-way
catalyst 511 in the upstream catalytic converter, and measures the
oxygen concentration of the exhaust gas;
Water temperature sensor SW10: Attached to the engine 1 and
measures the temperature of coolant;
Crank angle sensor SW11: Attached to the engine 1 and measures the
rotation angle of the crankshaft 15;
Accelerator opening sensor SW12: Attached to an accelerator pedal
mechanism and measures the accelerator opening corresponding to an
operating amount of the accelerator pedal;
Intake cam angle sensor SW13: Attached to the engine 1 and measures
the rotation angle of an intake cam shaft;
Exhaust cam angle sensor SW14: Attached to the engine 1 and
measures the rotation angle of an exhaust cam shaft;
EGR pressure difference sensor SW15: Disposed in the EGR passage 52
and measures a pressure difference between the upstream and the
downstream of the EGR valve 54;
Fuel pressure sensor SW16: Attached to the common rail 64 of the
fuel supply system 61, and measures the pressure of fuel supplied
to the injector 6; and
Third intake-air temperature sensor SW17: Attached to the surge
tank 42, and measures the temperature of gas inside the surge tank
42, i.e., the temperature of intake air introduced into the
combustion chamber 17.
The ECU 10 determines the operating state of the engine 1 based on
the signals of the sensors SW1-SW17, and calculates a control
amount of each device according to the control logic defined
beforehand. The control logic is stored in the memory 102. The
control logic includes calculating a target amount and/or the
control amount by using a map stored in the memory 102.
The ECU 10 outputs electrical signals according to the calculated
control amounts to the injectors 6, the ignition plugs 25, the
intake-side electric S-VT 23, the exhaust-side electric S-VT 24,
the fuel supply system 61, the throttle valve 43, the EGR valve 54,
the electromagnetic clutch 45 of the supercharger 44, the air
bypass valve 48, and the swirl control valve 56.
For example, the ECU 10 sets a target torque of the engine 1 based
on the signal of the accelerator opening sensor SW12 and the map,
and determines a target supercharging pressure. The ECU 10 then
performs a feedback control for adjusting the opening of the air
bypass valve 48 based on the target supercharging pressure and the
pressure difference before and after the supercharger 44 obtained
from the signals of the first pressure sensor SW3 and the second
pressure sensor SW5 so that the supercharging pressure becomes the
target supercharging pressure.
Moreover, the ECU 10 sets a target EGR ratio (i.e., a ratio of the
EGR gas to the entire gas inside the combustion chamber 17) based
on the operating state of the engine 1 and the map. The ECU 10 then
determines a target EGR gas amount based on the target EGR ratio
and an intake air amount based on the signal of the accelerator
opening sensor SW12, and performs a feedback control for adjusting
the opening of the EGR valve 54 based on the pressure difference
before and after the EGR valve 54 obtained from the signal of the
EGR pressure difference sensor SW15 so that the external EGR gas
amount introduced into the combustion chamber 17 becomes the target
EGR gas amount.
Further, the ECU 10 performs an air-fuel ratio feedback control
when a given control condition is satisfied. For example, the ECU
10 adjusts the fuel injection amount of the injector 6 based on the
oxygen concentration of the exhaust gas which is measured by the
linear O.sub.2 sensor SW8 and the lambda O.sub.2 sensor SW9 so that
the air-fuel ratio of the mixture gas becomes a desired value.
Note that the details of other controls of the engine 1 executed by
the ECU 10 will be described later.
(Concept of SPCCI Combustion)
The engine 1 performs combustion by compressed self-ignition under
a given operating state, mainly to improve fuel consumption and
emission performance. The combustion by self-ignition varies
largely at the timing of the self-ignition, if the temperature
inside the combustion chamber 17 before a compression starts is
nonuniform. Thus, the engine 1 performs the SPCCI combustion which
is a combination of the SI combustion and the CI combustion.
The SPCCI combustion is combustion in which the ignition plug 25
forcibly ignites the mixture gas inside the combustion chamber 17
so that the mixture gas carries out the SI combustion by flame
propagation, and the temperature inside the combustion chamber 17
increases by the heat generation of the SI combustion and the
pressure inside the combustion chamber 17 increases by the flame
propagation so that the unburnt mixture gas carries out the CI
combustion by self-ignition.
By adjusting the heat amount of the SI combustion, the variation in
the temperature inside the combustion chamber 17 before a
compression starts can be absorbed. By the ECU 10 adjusting the
ignition timing, the mixture gas can be self-ignited at a target
timing.
In the SPCCI combustion, the heat release of the SI combustion is
slower than the heat release in the CI combustion. As illustrated
in FIG. 5, the waveform of the heat release rate of the SI
combustion in the SPCCI combustion is smaller in the rising slope
than the waveform in the CI combustion. In addition, the SI
combustion is slower in the pressure fluctuation (dp/d.theta.)
inside the combustion chamber 17 than the CI combustion.
When the unburnt mixture gas self-ignites after the SI combustion
is started, the waveform slope of the heat release rate may become
steeper. The waveform of the heat release rate may have an
inflection point X at a timing of starting the CI combustion.
After the start in the CI combustion, the SI combustion and the CI
combustion are performed in parallel. Since the CI combustion has
larger heat release than the SI combustion, the heat release rate
becomes relatively large. However, since the CI combustion is
performed after a compression top dead center, the waveform slope
of the heat release rate does not become too steep. The pressure
fluctuation in the CI combustion (dp/d.theta.) also becomes
comparatively slow.
The pressure fluctuation (dp/d.theta.) can be used as an index
representing combustion noise. As described above, since the SPCCI
combustion can reduce the pressure fluctuation (dp/d.theta.), it is
possible to avoid excessive combustion noise. Combustion noise of
the engine 1 can be kept below the tolerable level.
The SPCCI combustion is completed when the CI combustion is
finished. The CI combustion is shorter in the combustion period
than the SI combustion. The end timing of the SPCCI combustion
becomes earlier than the SI combustion.
The heat release rate waveform of the SPCCI combustion is formed so
that a first heat release rate part Q.sub.SI formed by the SI
combustion and a second heat release rate part Q.sub.CI formed by
the CI combustion continue in this order.
Here, a SI ratio is defined as a parameter indicative of a
characteristic of the SPCCI combustion. The SI ratio is defined as
an index related to a ratio of the amount of heat generated by the
SI combustion to the entire amount of heat generated by the SPCCI
combustion. The SI ratio is a ratio of heat amount generated by the
two different combustion modes. If the SI ratio is high, the ratio
of the SI combustion is high, and if the SI ratio is low, the ratio
in the CI combustion is high. Alternatively, the SI ratio may be
defined as a ratio of the amount of heat generated by the SI
combustion to the amount of heat generated by the CI combustion.
That is, in a waveform 801 illustrated in FIG. 5, SI
ratio=Q.sub.SI/Q.sub.CI. Here,
Q.sub.SI: Area of SI combustion; and
Q.sub.CI: Area of CI combustion.
The engine 1 generates a strong swirl flow inside the combustion
chamber 17 when performing the SPCCI combustion. The term "strong
swirl flow" may be defined as a flow having a swirl ratio of four
or higher, for example. The swirl ratio can be defined as a value
obtained by dividing an integrated value of intake flow lateral
angular velocities by an engine angular velocity, where the intake
flow lateral angular velocity is measured for every valve lift, and
the measured values are integrated to obtain the integrated value.
Although illustration is omitted, the intake flow lateral angular
velocity can be obtained based on measurement using known rig test
equipment.
When the strong swirl flow is generated in the combustion chamber
17, the swirl flow is stronger in an outer circumferential part of
the combustion chamber 17 and is relatively weaker in a central
part. By the whirlpool resulting from a velocity gradient at the
boundary between the central part and the outer circumferential
part, turbulence energy becomes higher in the central part. When
the ignition plug 25 ignites the mixture gas in the central part,
the combustion speed of the SI combustion becomes higher by the
high turbulence energy.
Flame of the SI combustion is carried by the strong swirl flow
inside the combustion chamber 17 and propagates in the
circumferential direction. The CI combustion is performed from the
outer circumferential part to the central part in the combustion
chamber 17.
When the strong swirl flow is generated in the combustion chamber
17, the SI combustion can fully be performed before the start in
the CI combustion. Thus, the generation of combustion noise can be
reduced, and the variation in the torque between cycles can be
reduced.
(Engine Operating Range)
FIGS. 6 and 7 illustrate maps according to the control of the
engine 1. The maps are stored in the memory 102 of the ECU 10. The
maps include three kinds of maps, a map 501, a map 502, and a map
503. The ECU 10 uses a map selected from the three kinds of maps
501, 502, and 503 according to a wall temperature of the combustion
chamber 17 and an intake air temperature, in order to control the
engine 1. Note that the details of the selection of the three kinds
of maps 501, 502, and 503 will be described later.
The first map 501 is a map when the engine 1 is warm. The second
map 502 is a map when the engine 1 is half warm. The third map 503
is a map when the engine 1 is cold.
The maps 501, 502, and 503 are defined based on the load and the
engine speed of the engine 1. The first map 501 is roughly divided
into three areas depending on the load and the engine speed. For
example, the three areas include a low load area A1, a
middle-to-high load area (A2, A3, and A4), and a high speed area
A5. The low load area A1 includes idle operation, and covers areas
of a low engine speed and a middle engine speed. The middle-to-high
load area (A2, A3, and A4) are higher in the load than the low load
area A1. The high speed area A5 is higher in the engine speed than
the low load area A1 and the middle-to-high load area (A2, A3, and
A4). The middle-to-high load area (A2, A3, and A4) is divided into
a middle load area A2, a high-load middle-speed area A3 where the
load is higher than the middle load area A2, and a high-load
low-speed area A4 where the engine speed is lower than the
high-load middle-speed area A3.
The second map 502 is roughly divided into two areas. For example,
the two areas include a low-to-middle speed area (B1, B2, and B3)
and a high speed area B4 where the engine speed is higher than the
low-to-middle speed area (B1, B2, and B3). The low-to-middle speed
area (B1, B2, and B3) is divided into a low-to-middle load area B1
corresponding to the low load area A1 and the middle load area A2,
a high-load middle-speed area B2, and a high-load low-speed area
B3.
The third map 503 has only one area Cl, without being divided into
a plurality of areas.
Here, the low speed area, the middle speed area, and the high speed
area may be defined by substantially equally dividing the entire
operating range of the engine 1 into three areas in the engine
speed direction. In the example of FIGS. 6 and 7, the engine speed
is defined to be a low speed if the engine speed is lower than the
engine speed N1, a high speed if the engine speed is higher than or
equal to the engine speed N2, and a middle speed if the engine
speed is higher than or equal to the engine speed N1 and lower than
the engine speed N2. For example, the engine speed N1 may be about
1,200 rpm, and the engine speed N2 may be about 4,000 rpm.
Moreover, the low load area may be an area including an operating
state with the light load, the high load area may be an area
including an operating state with full load, and the middle load
area may be an area between the low load area and the high load
area. Moreover, the low load area, the middle load area, and the
high load area may be defined by substantially equally dividing the
entire operating range of the engine 1 into three areas in the load
direction.
The maps 501, 502, and 503 in FIG. 6 illustrate the states and
combustion modes of the mixture gas in the respective areas. A map
504 in FIG. 7 corresponds to the first map 501, and illustrates the
state and combustion mode of the mixture gas in each area of the
map, the opening of the swirl control valve 56 in each area, and a
driving area and a non-driving area of the supercharger 44. The
engine 1 performs the SPCCI combustion in the low load area A1, the
middle load area A2, the high-load middle-speed area A3, the
high-load low-speed area A4, the low-to-middle load area B1, the
high-load middle-speed area B2, and the high-load low-speed area
B3. The engine 1 performs the SI combustion in other areas,
specifically, in the high speed area A5, the high speed area B4,
and the area C1.
(Operation of Engine in Each Area)
Below, the operation of the engine 1 in each area of the map 504 in
FIG. 7 will be described in detail with reference to the fuel
injection timing and the ignition timing which are illustrated in
FIG. 8. The horizontal axis in FIG. 8 is a crank angle. Note that
the reference numerals 601, 602, 603, 604, 605, and 606 in FIG. 8
correspond to the operating states of the engine 1 indicated by the
reference numerals 601, 602, 603, 604, 605, and 606 in the map 504
of FIG. 7, respectively.
(Low load Area)
The engine 1 performs the SPCCI combustion when the engine 1
operates in the low load area A1.
The reference numeral 601 in FIG. 8 indicates fuel injection
timings (reference numerals 6011 and 6012), an ignition timing
(reference numeral 6013), and a combustion waveform (i.e., a
waveform indicating a change in the heat release rate with respect
to the crank angle: reference numeral 6014), when the engine 1
operates in the operating state 601 in the low load area A1. The
reference numeral 602 indicates fuel injection timings (reference
numerals 6021 and 6022), an ignition timing (reference numeral
6023), and a combustion waveform (reference numeral 6024), when the
engine 1 operates in the operating state 602 in the low load area
A1. The reference numeral 603 indicates fuel injection timings
(reference numerals 6031 and 6032), an ignition timing (reference
numeral 6033), and a combustion waveform (reference numeral 6034),
when the engine 1 operates in the operating state 603 in the low
load area A1. The operating states 601, 602, and 603 have the same
engine speed, but different loads. The operating state 601 has the
lowest load (i.e., light load), the operating state 602 has the
second lowest load (i.e., low load), and the operating state 603
has the maximum load among these states.
In order to improve the fuel efficiency of the engine 1, the EGR
system 55 introduces the EGR gas into the combustion chamber 17.
For example, the intake-side electric S-VT 23 and the exhaust-side
electric S-VT 24 are provided with a positive overlap period where
both the intake valve 21 and the exhaust valve 22 are opened near
an exhaust top dead center. Part of the exhaust gas discharged from
the combustion chamber 17 into the intake port 18 and the exhaust
port 19 is re-introduced into the combustion chamber 17. Since the
hot exhaust gas is introduced into the combustion chamber 17, the
temperature inside the combustion chamber 17 increases. Thus, it
becomes advantageous to stabilize the SPCCI combustion. Note that
the intake-side electric S-VT 23 and the exhaust-side electric S-VT
24 may be provided with a negative overlap period where both the
intake valve 21 and the exhaust valve 22 are closed.
Moreover, the swirl generating part forms the strong swirl flow
inside the combustion chamber 17. The swirl ratio is four or
higher, for example. The swirl control valve 56 is fully closed or
at a given opening (closed to some extent). As described above,
since the intake port 18 is the tumble port, an inclined swirl flow
having a tumble component and a swirl component is formed in the
combustion chamber 17.
The injector 6 injects fuel into the combustion chamber 17 a
plurality of times during the intake stroke (reference numerals
6011, 6012, 6021, 6022, 6031, and 6032). The mixture gas is
stratified by the multiple fuel injections and the swirl flow
inside the combustion chamber 17.
The fuel concentration of the mixture gas in the central part of
the combustion chamber 17 is denser or richer than the fuel
concentration in the outer circumferential part. For example, the
air-fuel ratio (A/F) of the mixture gas in the central part is 20
or higher and 30 or lower, and the A/F of the mixture gas in the
outer circumferential part is 35 or higher. Note that the value of
the A/F is a value when the mixture gas is ignited, and the same
applies to the following description. Since the A/F of the mixture
gas near the ignition plug 25 is set 20 or higher and 30 or lower,
generation of raw NO.sub.x during the SI combustion can be reduced.
Moreover, since the A/F of the mixture gas in the outer
circumferential part is set to 35 or higher, the CI combustion
stabilizes.
The A/F of the mixture gas is leaner than the stoichiometric air
fuel ratio throughout the combustion chamber 17 (i.e., excess air
ratio .lamda.>1). For example, the A/F of the mixture gas is 30
or higher throughout the combustion chamber 17. Thus, the
generation of raw NO.sub.x can be reduced to improve the emission
performance.
When the engine load is low (i.e., in the operating state 601), the
injector 6 performs the first injection 6011 in the first half of
an intake stroke, and performs the second injection 6012 in the
second half of the intake stroke. The first half of the intake
stroke may be a first half of an intake stroke when the intake
stroke is equally divided into two, and the second half of the
intake stroke may be the rest. Moreover, an injection amount ratio
of the first injection 6011 to the second injection 6012 may be
9:1, for example.
In the operating state 602 where the engine load is higher, the
injector 6 initiates the second injection 6022 which is performed
in the second half of an intake stroke at a timing advanced from
the second injection 6012 in the operating state 601. By advancing
the second injection 6022, the mixture gas inside the combustion
chamber 17 becomes more homogeneous. The injection amount ratio of
the first injection 6021 to the second injection 6022 may be 7:3 to
8:2, for example.
In the operating state 603 where the engine load is even higher,
the injector 6 initiates the second injection 6032 which is
performed in the second half of an intake stroke at a timing
further advanced from the second injection 6022 in the operating
state 602. By further advancing the second injection 6032, the
mixture gas inside the combustion chamber 17 becomes further
homogeneous. The injection amount ratio of the first injection 6031
to the second injection 6032 may be 6:4, for example.
After the fuel injection is finished, the ignition plug 25 ignites
the mixture gas in the central part of the combustion chamber 17 at
a given timing before a compression top dead center (reference
numerals 6013, 6023, and 6033). The ignition timing may be during a
final stage of the compression stroke. The compression stroke may
be equally divided into three, an initial stage, a middle stage,
and a final stage, and this finale stage may be used as the final
stage of the compression stroke described above.
As described above, since the mixture gas in the central part has
the relatively high fuel concentration, the ignitability improves
and the SI combustion by flame propagation stabilizes. By the SI
combustion being stabilized, the CI combustion begins at a suitable
timing. Thus, the controllability in the CI combustion improves in
the SPCCI combustion. Further, the generation of combustion noise
is reduced. Moreover, since the A/F of the mixture gas is made
leaner than the stoichiometric air fuel ratio to perform the SPCCI
combustion, fuel efficiency of the engine 1 can be significantly
improved. Note that the low load area A1 corresponds to Layer 3
described later. Layer 3 extends to the light load operating range
and includes a minimum load operating state.
(Middle-to-High Load Area)
When the engine 1 operates in the middle-to-high load area, the
engine 1 also performs the SPCCI combustion, similar to the low
load area.
The reference numeral 604 in FIG. 8 indicates, in the
middle-to-high load area, fuel injection timings (reference
numerals 6041 and 6042), an ignition timing (reference numeral
6043), and a combustion waveform (reference numeral 6044), when the
engine 1 operates in the operating state 604 in the middle load
area A2. The reference numeral 605 indicates a fuel injection
timing (reference numeral 6051), an ignition timing (reference
numeral 6052), and a combustion waveform (reference numeral 6053),
when the engine 1 operates in the operating state 605 in the
high-load low-speed area A4.
The EGR system 55 introduces the EGR gas into the combustion
chamber 17. For example, the intake-side electric S-VT 23 and the
exhaust-side electric S-VT 24 are provided with a positive overlap
period where both the intake valve 21 and the exhaust valve 22 are
opened near an exhaust top dead center. Internal EGR gas is
introduced into the combustion chamber 17. Moreover, the EGR system
55 introduces the exhaust gas cooled by the EGR cooler 53 into the
combustion chamber 17 through the EGR passage 52. That is, the
external EGR gas with a lower temperature than the internal EGR gas
is introduced into the combustion chamber 17. The external EGR gas
adjusts the temperature inside the combustion chamber 17 to a
suitable temperature. The EGR system 55 reduces the amount of the
EGR gas as the engine load increases. The EGR system 55 may not
recirculate the EGR gas containing the internal EGR gas and the
external EGR gas during the full load.
Moreover, in the middle load area A2 and the high-load middle-speed
area A3, the swirl control valve 56 is fully closed or at a given
opening (closed to some extent). In the combustion chamber 17, the
strong swirl flow with the swirl ratio of four or higher is formed.
On the other hand, in the high-load low-speed area A4, the swirl
control valve 56 is open.
The air-fuel ratio (A/F) of the mixture gas is the stoichiometric
air fuel ratio (A/F.apprxeq.14.7:1) throughout the combustion
chamber 17. Since the three-way catalysts 511 and 513 purify the
exhaust gas discharged from the combustion chamber 17, the emission
performance of the engine 1 is improved. The A/F of the mixture gas
may be set within a purification window of the three-way catalyst.
The excess air ratio .lamda. of the mixture gas may be 1.0.+-.0.2.
Note that when the engine 1 operates in the high-load middle-speed
area A3 including the full load (i.e., the maximum load), the A/F
of the mixture gas may be set at the stoichiometric air fuel ratio
or richer than the stoichiometric air fuel ratio (i.e., the excess
air ratio .lamda. of the mixture gas is .lamda..ltoreq.1)
throughout the combustion chamber 17.
Since the EGR gas is introduced into the combustion chamber 17, the
G/F which is a weight ratio of the entire gas to the fuel in the
combustion chamber 17 becomes leaner than the stoichiometric air
fuel ratio. The G/F of the mixture gas may be 18:1 or higher. Thus,
a generation of a so-called "knock" is avoided. The G/F may be set
18:1 or higher and 30:1 or lower. Alternatively, the G/F may be set
18:1 or higher and 50:1 or lower.
When the engine 1 operates in the operating state 604, the injector
6 performs a plurality of fuel injections (reference numerals 6041
and 6042) during an intake stroke. The injector 6 may perform the
first injection 6041 in the first half of the intake stroke and the
second injection 6042 in the second half of the intake stroke.
Moreover, when the engine 1 operates in the operating state 605,
the injector 6 injects fuel in an intake stroke (reference numeral
6051).
The ignition plug 25 ignites the mixture gas at a given timing near
a compression top dead center after the fuel is injected (reference
numerals 6043 and 6052). The ignition plug 25 may ignite the
mixture gas before the compression top dead center when the engine
1 operates in the operating state 604 (reference numeral 6043). The
ignition plug 25 may ignite the mixture gas after the compression
top dead center when the engine 1 operates in the operating state
605 (reference numeral 6052).
Since the A/F of the mixture gas is set to the stoichiometric air
fuel ratio and the SPCCI combustion is performed, the exhaust gas
discharged from the combustion chamber 17 can be purified using the
three-way catalysts 511 and 513. Moreover, the fuel efficiency of
the engine 1 improves by introducing the EGR gas into the
combustion chamber 17 and making the mixture gas leaner. Note that
the middle-to-high load areas A2, A3, and A4 correspond to Layer 2
described later. Layer 2 extends to the high load area and includes
the maximum load operating state.
(Operation of Supercharger)
Here, as illustrated in the map 504 of FIG. 7, the supercharger 44
is OFF in part of the low load area A1 and part of the middle load
area A2 (see S/C OFF). In detail, the supercharger 44 is OFF in an
area on the lower engine speed side in the low load area A1. In an
area on the higher speed side in the low load area A1, the
supercharger 44 is ON in order to secure a required intake filling
amount for the increased speed of the engine 1. Moreover, the
supercharger 44 is OFF in a partial area on the lower load and
lower engine speed side in the middle load area A2. In the area on
the higher load side in the middle load area A2, the supercharger
44 is ON in order to secure a required intake filling amount for
the increased fuel injection amount. Moreover, the supercharger 44
is ON also in the area on the higher speed side in the middle load
area A2.
Note that in each area of the high-load middle-speed area A3, the
high-load low-speed area A4, and the high speed area A5, the
supercharger 44 is entirely ON (see S/C ON).
(High Speed Area)
When the speed of the engine 1 increases, a time required for
changing the crank angle by 1.degree. becomes shorter. Thus, it
becomes difficult to stratify the mixture gas inside the combustion
chamber 17. When the speed of the engine 1 increases, it also
becomes difficult to perform the SPCCI combustion.
Therefore, the engine 1 performs not the SPCCI combustion but the
SI combustion when the engine 1 operates in the high speed area A5.
Note that the high speed area A5 extends entirely in the load
direction from the low load to the high load.
The reference numeral 606 of FIG. 8 indicates a fuel injection
timing (reference numeral 6061), an ignition timing (reference
numeral 6062), and a combustion waveform (reference numeral 6063),
when the engine 1 operates in the high speed area A5 in the
operating state 606 where the load is high.
The EGR system 55 introduces the EGR gas into the combustion
chamber 17. The EGR system 55 reduces the amount of the EGR gas as
the load increases. The EGR system 55 may not recirculate the EGR
gas during full load.
The swirl control valve 56 is fully opened. No swirl flow occurs in
the combustion chamber 17, but only a tumble flow occurs. By fully
opening the swirl control valve 56, it is possible to increase the
filling efficiency, and reduce the pumping loss.
Fundamentally, the air-fuel ratio (A/F) of the mixture gas is the
stoichiometric air fuel ratio (A/F.apprxeq.14.7:1) throughout the
combustion chamber 17. The excess air ratio .lamda. of the mixture
gas may be set to 1.0.+-.0.2. Note that the excess air ratio
.lamda. of the mixture gas may be lower than 1 when the engine 1
operates near full load.
The injector 6 starts the fuel injection during an intake stroke.
The injector 6 injects the fuel all at once (reference numeral
6061). By starting the fuel injection in the intake stroke, the
homogeneous or substantially homogeneous mixture gas is formed
inside the combustion chamber 17. Moreover, since a longer
vaporizing time of the fuel can be secured, unburnt fuel loss can
also be reduced.
After the fuel injection is finished, the ignition plug 25 ignites
the mixture gas at a suitable timing before a compression top dead
center (reference numeral 6062).
(Layer Structure of Map)
As illustrated in FIG. 9, the maps 501, 502, and 503 of the engine
1 illustrated in FIG. 6 are comprised of a combination of three
layers, Layer 1, Layer 2, and Layer 3.
Layer 1 is a layer used as a base layer. Layer 1 extends throughout
the operating range of the engine 1. Layer 1 corresponds to the
entire third map 503.
Layer 2 is a layer which is superimposed on Layer 1. Layer 2
corresponds to part of the operating range of the engine 1. For
example, Layer 2 corresponds to the low-to-middle speed area B1,
B2, and B3 of the second map 502.
Layer 3 is a layer which is superimposed on Layer 2. Layer 3
corresponds to the low load area A1 of the first map 501.
Layer 1, Layer 2, and Layer 3 are selected according to the wall
temperature of the combustion chamber 17 and the intake air
temperature.
When the wall temperature of the combustion chamber 17 is higher
than a given first wall temperature (e.g., 80.degree. C.) and the
intake air temperature is higher than a given first intake air
temperature (e.g., 50.degree. C.), Layer 1, Layer 2, and Layer 3
are selected, and the first map 501 is formed by piling up Layer 1,
Layer 2, and Layer 3. In the low load area A1 in the first map 501,
the top Layer 3 therein becomes effective, in the middle-to-high
load areas A2, A3, and A4, the top Layer 2 therein becomes
effective, and in the high speed area A5, Layer 1 becomes
effective.
When the wall temperature of the combustion chamber 17 is lower
than the given first wall temperature and higher than a given
second wall temperature (e.g., 30.degree. C.), and the intake air
temperature is lower than the given first intake air temperature
and higher than a given second intake air temperature (e.g.,
25.degree. C.), Layer 1 and Layer 2 are selected. By superimposing
the Layer 1 and Layer 2, the second map 502 is formed. The
low-to-middle speed area B1, B2, and B3 in second map 502, the top
Layer 2 therein becomes effective, and in the high speed area B4,
Layer 1 becomes effective.
When the wall temperature of the combustion chamber 17 is lower
than the given second wall temperature and the intake air
temperature is lower than the given second intake air temperature,
only Layer 1 is selected to form the third map 503.
Note that the wall temperature of the combustion chamber 17 may be
replaced, for example, by temperature of the coolant of the engine
1 measured by the water temperature sensor SW10. Alternatively, the
wall temperature of the combustion chamber 17 may be estimated
based on the temperature of the coolant, or other measurements. The
intake air temperature is measurable, for example, by the third
intake-air temperature sensor SW17 which measures the temperature
inside the surge tank 42. Alternatively, the temperature of the
intake air introduced into the combustion chamber 17 may be
estimated based on various kinds of measurements.
As described above, the SPCCI combustion is performed by generating
the strong swirl flow inside the combustion chamber 17. Since the
flame propagates along the wall of the combustion chamber 17 during
the SI combustion, the flame propagation of the SI combustion is
influenced by the wall temperature. If the wall temperature is low,
the flame of the SI combustion is cooled to delay the timing of
compression-ignition.
Since the CI combustion of the SPCCI combustion is performed in the
area from the outer circumferential part to the central part of the
combustion chamber 17, it is influenced by the temperature in the
central part of the combustion chamber 17. If the temperature in
the central part is low, the CI combustion becomes unstable. The
temperature in the central part of the combustion chamber 17
depends on the temperature of the intake air introduced into the
combustion chamber 17. That is, when the intake air temperature is
higher, the temperature in the central part of the combustion
chamber 17 becomes higher, and when the intake air temperature is
lower, the temperature in the central part becomes lower.
When the wall temperature of the combustion chamber 17 is lower
than the given second wall temperature and the intake air
temperature is lower than the given second intake air temperature,
a stable SPCCI combustion cannot be performed. Thus, only Layer 1
which performs the SI combustion is selected, and the ECU 10
operates the engine 1 based on the third map 503. By the engine 1
performing the SI combustion in the entire operating range, the
combustion stability can be secured.
When the wall temperature of the combustion chamber 17 is higher
than the given second wall temperature and the intake air
temperature is higher than the given second intake air temperature,
the stable SPCCI combustion of the mixture gas having substantially
stoichiometric air fuel ratio (i.e., .lamda.=1) can be carried out.
Thus, in addition to Layer 1, Layer 2 is selected, and the ECU 10
operates the engine 1 based on the second map 502. By the engine 1
performing the SPCCI combustion in the part of the operating
ranges, the fuel efficiency of the engine 1 improves.
When the wall temperature of the combustion chamber 17 is higher
than the given first wall temperature and the intake air
temperature is higher than the given first the intake air
temperature, the stable SPCCI combustion of the mixture gas leaner
than the stoichiometric air fuel ratio can be carried out. Thus, in
addition to Layer 1 and Layer 2, Layer 3 is selected, and the ECU
10 operates the engine 1 based on the first map 501. By the engine
1 performing the SPCCI combustion of the lean mixture gas in the
part of the operating ranges, the fuel efficiency of the engine 1
further improves.
Next, one example of control related to the layer selection of the
map executed by the ECU 10 will be described with reference to a
flowchart of FIG. 10. First, at Step S1 after the control is
started, the ECU 10 reads the signals of the sensors SW1-SW17. At
the following Step S2, the ECU 10 determines whether the wall
temperature of the combustion chamber 17 is 30.degree. C. or higher
and the intake air temperature is 25.degree. C. or higher. If the
determination at Step S2 is YES, the control shifts the process to
Step S3, and on the other hand, if NO, the control shifts the
process to Step S5. The ECU 10 selects only Layer 1 at Step S5. The
ECU 10 operates the engine 1 based on the third map 503. The
control then returns the process.
At Step S3, the ECU 10 determines whether the wall temperature of
the combustion chamber 17 is 80.degree. C. or higher and the intake
air temperature is 50.degree. C. or higher. If the determination at
Step S3 is YES, the control shifts the process to Step S4, and on
the other hand, if NO, the control shifts the process to Step
S6.
The ECU 10 selects Layer 1 and Layer 2 at Step S6. The ECU 10
operates the engine 1 based on the second map 502. The control then
returns the process.
The ECU 10 selects Layer 1, Layer 2, and Layer 3 at Step S4. The
ECU 10 operates the engine 1 based on the first map 501. The
control then returns the process.
(Valve Timing of Intake Valve and Exhaust Valve)
FIG. 11 illustrates one example of a change in an open timing IVO
of the intake valve 21 when the ECU 10 controls the intake-side
electric S-VT 23 according to the control logic set for Layer 2. An
upper figure of FIG. 11 (i.e., a graph 1101) illustrates a change
of the open timing IVO of the intake valve 21 (vertical axis)
versus the engine load (horizontal axis). The solid line
corresponds to a case where the speed of the engine 1 is a
relatively low first engine speed, and a broken line corresponds to
a case where the speed of the engine 1 is a relatively high second
engine speed (first engine speed<second engine speed).
A lower figure of FIG. 11 (i.e., a graph 1102) illustrates a change
of the open timing IVO of the intake valve 21 (vertical axis)
versus the speed of the engine 1 (horizontal axis). The solid line
corresponds to a case where the engine load is a relatively low
first load, and the broken line corresponds to a case where the
engine load is a relatively high second load (first load <second
load).
In the graph 1101 and the graph 1102, the open timing IVO of the
intake valve 21 is advanced as it goes upward and the positive
overlap period where both the intake valve 21 and the exhaust valve
22 open becomes longer. Therefore, the amount of the EGR gas
introduced into the combustion chamber 17 increases.
In Layer 2, the engine 1 operates with the A/F of the mixture gas
at the stoichiometric air fuel ratio or the substantially
stoichiometric air fuel ratio, and the G/F leaner than the
stoichiometric air fuel ratio. When the engine load is low, the
fuel supply amount decreases. As illustrated in the graph 1101,
when the engine load is low, the ECU 10 sets the open timing IVO of
the intake valve 21 at a timing on the retard side. Thus, the
amount of the EGR gas introduced into the combustion chamber 17 is
regulated to secure the combustion stability.
Since the fuel supply amount increases when the engine load
increases, the combustion stability improves. The ECU 10 sets the
open timing of the intake valve 21 at a timing on the advance side.
The pumping loss of the engine 1 can be lowered by increasing the
amount of the EGR gas introduced into the combustion chamber
17.
When the engine load further increases, the temperature inside the
combustion chamber 17 further increases. Then, the amount of the
internal EGR gas is reduced and the amount of the external EGR gas
is increased so that the temperature inside the combustion chamber
17 does not become too high. Therefore, the ECU 10 sets the open
timing of the intake valve 21 again at a timing on the retard
side.
When the engine load further increases and the supercharger 44
starts boosting, the ECU 10 sets the open timing of the intake
valve 21 again at a timing on the advance side. Since the positive
overlap period where both the intake valve 21 and the exhaust valve
22 open is provided, the residual gas in the combustion chamber 17
can be purged.
Note that when the engine speed is high and low, the tendency of
the change in the open timing of the intake valve 21 is almost the
same.
As illustrated in the graph 1102, when the engine speed is low, the
flow inside the combustion chamber 17 becomes weaker. Since the
combustion stability falls, the amount of the EGR gas introduced
into the combustion chamber 17 is regulated. The ECU 10 sets the
open timing of the intake valve 21 at a timing on the retard
side.
Since the flow inside the combustion chamber 17 becomes strong when
the engine speed increases, the amount of the EGR gas introduced
into the combustion chamber 17 can be increased. The ECU 10 sets
the open timing of the intake valve 21 at a timing on the advance
side.
When the engine speed further increases, the ECU 10 sets the open
timing of the intake valve 21 at a timing on the retard side
according to the engine speed. Thus, the amount of gas introduced
into the combustion chamber 17 is maximized.
FIG. 12 illustrates one example of a change in the open timing IVO
of the intake valve 21, a close timing EVC of the exhaust valve 22,
and an overlap period O/L of the intake valve 21 and the exhaust
valve 22, when the ECU 10 controls the intake-side electric S-VT 23
and the exhaust-side electric S-VT 24 according to the control
logic set for Layer 3.
An upper figure of FIG. 12 (i.e., a graph 1201) illustrates a
change in the open timing IVO of the intake valve 21 (vertical
axis) versus the engine load (horizontal axis). The solid line
corresponds to a case where the engine speed is a relatively low
third engine speed, and the broken line corresponds to a case where
the engine speed is relatively high fourth engine speed (third
engine speed<fourth engine speed).
A middle figure of FIG. 12 (i.e., graph 1202) illustrates a change
in the close timing EVC of the exhaust valve 22 (vertical axis)
versus the engine load (horizontal axis). The solid line
corresponds to a case where the engine speed is at the third engine
speed, and the broken line corresponds to a case where the engine
speed is at the fourth engine speed.
A lower figure of FIG. 12 (i.e., graph 1203) illustrates a change
in the overlap period O/L of the intake valve 21 and the exhaust
valve 22 (vertical axis) versus the engine load (horizontal axis).
The solid line corresponds to a case where the engine speed is at
the third engine speed, and the broken line corresponds to a case
where the engine speed is at the fourth engine speed.
In Layer 3, the engine 1 operates by carrying out the SPCCI
combustion of the mixture gas with the A/F leaner than the
stoichiometric air fuel ratio. When the engine load is low, the
fuel supply amount decreases. Thus, the ECU 10 regulates the amount
of gas introduced into the combustion chamber 17 so that the A/F of
the mixture gas does not become too low. As illustrated in the
graph 1201, the ECU 10 sets the open timing IVO of the intake valve
21 at a timing on the retard side of an exhaust top dead center.
The close timing of the intake valve 21 becomes after an intake
bottom dead center, so-called "late close". Moreover, when the
engine load is low, the ECU 10 regulates the amount of the internal
EGR gas introduced into the combustion chamber 17. As illustrated
in the graph 1202, the ECU 10 sets the close timing EVC of the
exhaust valve 22 at a timing on the advance side. The close timing
EVC of the exhaust valve 22 approaches an exhaust top dead
center.
Since the fuel supply amount increases when the engine load
increases, the ECU 10 does not regulate the amount of gas
introduced into the combustion chamber 17. Moreover, in order to
stabilize the SPCCI combustion of the mixture gas leaner than the
stoichiometric air fuel ratio, the ECU 10 increases the amount of
the internal EGR gas introduced into the combustion chamber 17. The
ECU 10 sets the open timing IVO of the intake valve 21 at a timing
on the advance side of an exhaust top dead center. Moreover, the
ECU 10 sets the close timing EVC of the exhaust valve 22 at a
timing on the retard side of the exhaust top dead center. As a
result, as illustrated in the graph 1203, when the engine load
increases, the overlap period where both the intake valve 21 and
the exhaust valve 22 open becomes longer.
When the engine load further increases, the ECU 10 reduces the
amount of the internal EGR gas introduced into the combustion
chamber 17 so that the temperature inside the combustion chamber 17
does not become too high. The ECU 10 brings the close timing EVC of
the exhaust valve 22 closer to an exhaust top dead center. Thus,
the overlap period of the intake valve 21 and the exhaust valve 22
becomes shorter. Moreover, when the engine load is high and the
engine speed is high, the ECU 10 sets the open timing of the intake
valve 21 on the retard side more than when the engine speed is low.
Thus, the amount of gas introduced into the combustion chamber 17
is maximized.
Note that in the low load area surrounded by a one-dot chain line
in FIG. 12, the engine 1 may perform a reduced-cylinder operation
in order to improve fuel efficiency. When performing the
reduced-cylinder operation, the amount of gas and the internal EGR
gas introduced into the combustion chamber 17 are not regulated. As
illustrated by a two-dot chain line in the graphs 1201 and 1202,
the ECU 10 may set the open timing of the intake valve 21 at a
timing on the advance side and the close timing of the exhaust
valve 22 at a timing on the retard side.
(Engine Control Logic)
FIG. 13 is a flowchart illustrating the control logic of the engine
1. The ECU 10 operates the engine 1 according to the control logic
stored in the memory 102. For example, the ECU 10 determines the
operating state of the engine 1 based on the signals of the sensors
SW1-SW17, and performs calculations for adjusting properties in the
combustion chamber 17, the injection amount, the injection timing,
and the ignition timing so that the combustion in the combustion
chamber 17 becomes combustion at the SI ratio according to the
operating state.
The ECU 10 first reads the signals of the sensors SW1-SW17 at Step
S131. Subsequently, at Step S132, the ECU 10 determines the
operating state of the engine 1 based on the signals of the sensors
SW1-SW17, and sets a target SI ratio (i.e., a target heat amount
ratio). The target SI ratio is set according to the operating state
of the engine 1.
FIG. 14 schematically illustrates one example of a setting of the
target SI ratio. When the engine load is low, the ECU 10 sets the
target SI ratio low, and on the other hand, when the engine load is
high, it sets the target SI ratio high. When the engine load is
low, both the reduction of combustion noise and the improvement in
fuel efficiency can be achieved by increasing the ratio in the CI
combustion to the SPCCI combustion. When the engine load is high,
it becomes advantageous for the reduction of the combustion noise
by increasing the ratio of the SI combustion to the SPCCI
combustion.
Returning to the flowchart of FIG. 13, the ECU 10 sets the target
in-cylinder properties for achieving the target SI ratio setting
based on the combustion model stored in the memory 102 at the
following Step S133. For example, the ECU 10 sets a target
temperature, a target pressure, and target properties in the
combustion chamber 17. At Step S134, the ECU 10 sets an opening of
the EGR valve 54, an opening of the throttle valve 43, an opening
of the air bypass valve 48, an opening of the swirl control valve
56, and phase angles of the intake-side electric S-VT 23, and the
exhaust-side electric S-VT 24 (i.e., a valve timing of the intake
valve 21 and a valve timing of the exhaust valve 22), which are
required for achieving the target in-cylinder properties. The ECU
10 sets the control amounts of these devices based on the map
stored in the memory 102. The ECU 10 outputs signals to the EGR
valve 54, the throttle valve 43, the air bypass valve 48, the swirl
control valve (SCV) 56, the intake-side electric S-VT 23, and the
exhaust-side electric S-VT 24 based on the control amount setting.
By each device operating based on the signal of the ECU 10, the
properties in the combustion chamber 17 become the target
properties.
The ECU 10 further calculates predicted values or estimated values
of the properties in the combustion chamber 17 based on the control
amount setting of each device. The predicted property value is a
predicted value of the property in the combustion chamber 17 before
the intake valve 21 is closed. The predicted property value is used
for setting of the fuel injection amount during the intake stroke
as will be described later. The estimated property value is an
estimated value of the property in the combustion chamber 17 after
the intake valve 21 is closed. The estimated property value is used
for setting of the fuel injection amount during a compression
stroke, and setting of the ignition timing, as will be described
later. The estimated property value is also used for calculation of
a property error of comparison with an actual combustion state.
At Step S135, the ECU 10 sets a fuel injection amount in the intake
stroke based on the predicted property value. When performing a
divided injection during the intake stroke, the ECU 10 sets the
injection amount of each injection. Note that when fuel is not
injected in the intake stroke, the injection amount of fuel is
zero. At Step S136, the ECU 10 outputs a signal to the injector 6
so that the injector 6 injects fuel into the combustion chamber 17
at given injection timing(s).
At Step S137, the ECU 10 sets a fuel injection amount in a
compression stroke based on the estimated property value and the
injection result of the fuel in the intake stroke. Note that when
fuel is not injected in the compression stroke, the injection
amount of fuel is zero. At Step S138, the ECU 10 outputs a signal
to the injector 6 so that the injector 6 injects fuel into the
combustion chamber 17 at an injection timing based on the preset
map.
At Step S139, the ECU 10 sets an ignition timing based on the
estimated property value and the injection result of the fuel in
the compression stroke. At Step S1310, the ECU 10 outputs a signal
to the ignition plug 25 so that the ignition plug 25 ignites the
mixture gas inside the combustion chamber 17 at the set ignition
timing.
By the ignition plug 25 igniting the mixture gas, the SI combustion
or the SPCCI combustion is performed inside the combustion chamber
17. At Step S1311, the ECU 10 reads the change in the pressure
inside the combustion chamber 17 measured by the pressure
indicating sensor SW6, and based on the change, the ECU 10
determines a combustion state of the mixture gas inside the
combustion chamber 17. At Step S1312, the ECU 10 compares the
measurement result of the combustion state with the estimated
property values estimated at Step S134, and calculates an error
between the estimated properties value and the actual properties.
The calculated error is used for the estimation at Step S134 in the
subsequent cycles. The ECU 10 adjusts the openings of the throttle
valve 43, the EGR valve 54, the swirl control valve 56, and/or the
air bypass valve 48, and the phase angles of the intake-side
electric S-VT 23 and the exhaust-side electric S-VT 24 so as to
eliminate the property error. Thus, the amount of the fresh air and
the EGR gas introduced into the combustion chamber 17 are
adjusted.
If the ECU 10 estimates that the temperature inside the combustion
chamber 17 will be lower than the target temperature again based on
the estimated property values, it advances, at Step S138, the
injection timing in the compression stroke more than the injection
timing based on the map so that the advancing of the ignition
timing becomes possible. On the other hand, if ECU 10 estimates
that the temperature inside the combustion chamber 17 will be
higher than the target temperature based on the estimated property
values, it retards, at Step S138, the injection timing in the
compression stroke more than the injection timing based on the map
so that the retarding of the ignition timing becomes possible.
That is, if the temperature inside the combustion chamber 17 is
low, a self-ignition timing .theta..sub.CI (see FIG. 5) of the
unburnt mixture gas is delayed after the SI combustion begins by
jump-spark ignition, which deviates the SI ratio from the target SI
ratio. In this case, an increase in the unburnt fuel and a fall in
emission performance will be caused.
Thus, when the ECU 10 estimates that the temperature inside the
combustion chamber 17 will be lower than the target temperature, it
advances the injection timing, and advances the ignition timing at
Step S1310. Since the heat release which is sufficient for the SI
combustion becomes possible by an earlier start of the SI
combustion, it can prevent that the self-ignition timing
.theta..sub.CI of the unburnt mixture gas is delayed when the
temperature inside the combustion chamber 17 is low. As a result,
the SI ratio approaches the target SI ratio.
On the other hand, when the temperature inside the combustion
chamber 17 is high, the unburnt mixture gas will carry out the
self-ignition shortly after the SI combustion begins by the
jump-spark ignition, which deviates the SI ratio from the target SI
ratio. In this case, combustion noise increases.
Thus, when the ECU 10 estimates that the temperature inside the
combustion chamber 17 will be higher than the target temperature,
it retards the injection timing, and retards the ignition timing at
Step S1310. Since the start of the SI combustion is delayed, when
the temperature inside the combustion chamber 17 is high, the
self-ignition timing .theta..sub.CI of the unburnt mixture gas
becoming early can be prevented. As a result, the SI ratio
approaches the target SI ratio.
The control logic of the engine 1 is configured to adjust the SI
ratio by using a property setting device including the throttle
valve 43, the EGR valve 54, the air bypass valve 48, the swirl
control valve 56, the intake-side electric S-VT 23, and the
exhaust-side electric S-VT 24. By adjusting the properties in the
combustion chamber 17, a rough adjustment of the SI ratio is
possible. The control logic of the engine 1 is also configured to
adjust the SI ratio by adjusting the injection timing and the
ignition timing of the fuel. By adjusting the injection timing and
the ignition timing, a difference between the cylinders can be
corrected, and a fine adjustment of the self-ignition timing can be
performed, for example. By adjusting the SI ratio by two steps, the
engine 1 can achieve the target SPCCI combustion accurately
corresponding to the operating state.
The ECU 10 also outputs signals to at least the EGR system 55 and
the ignition plug 25 so that the actual SI ratio by combustion
becomes the target SI ratio. Moreover, as described above, when the
engine load is high, since the ECU 10 makes the target SI ratio
higher than when the load is low, it then outputs the signals to at
least the EGR system 55 and the ignition plug 25 when the engine
load is high so that the SI ratio becomes higher than when the load
is low.
Note that the control of the engine 1 executed by the ECU 10 is not
limited to the control logic based on the combustion model
described above.
(Method of Controlling Engine)
When controlling the engine 1 for performing the SPCCI combustion
described above, the parameter related to the control amount of
each device is set. For example, if the device is the ignition plug
25, ignition energy and the ignition timing corresponding to the
operating state of the engine 1 are set. If the device is the
intake-side electric S-VT 23, the valve timing of the intake valve
21 corresponding to the operating state of the engine 1 is set.
Below, the setting of the valve timing of the intake valve 21 in
the method of controlling the engine 1 will be described with
reference to the drawings.
(1) Controlling Method According to Close Timing of Intake
Valve
In the engine 1 for performing the SPCCI combustion, in order to
reduce combustion noise and achieve the stable SPCCI combustion,
the present inventors found out that it was necessary to adjust the
temperature inside the combustion chamber 17 to a suitable
temperature at a start timing in the CI combustion (.theta..sub.CI:
see FIG. 5). That is, when the temperature inside the combustion
chamber 17 is low, the ignitability in the CI combustion decreases.
When the temperature inside the combustion chamber 17 is high,
combustion noise increases.
The temperature inside the combustion chamber 17 at the start
timing .theta..sub.CI in the CI combustion is mainly related to the
effective compression ratio of the engine 1. The effective
compression ratio of the engine 1 is determined by the geometric
compression ratio .epsilon. and the close timing IVC of the intake
valve 21. In order to put the engine for performing the SPCCI
combustion in practical use, the present inventors newly found that
a suitable IVC range existed within a range of the geometric
compression ratio .epsilon. where the SPCCI combustion may occur.
The technology disclosed herein is novel in that a given relation
was found out between the geometric compression ratio .epsilon. and
the close timing IVC of the intake valve 21, where the given
relation is required in order to put in practical use the engine
which performs the unique combustion mode that is the SPCCI
combustion. In addition, a relational expression between the
geometric compression ratio .epsilon. and the close timing IVC of
the intake valve 21, which will be described later, is also
novel.
Further, the technology disclosed herein is also novel in that,
when controlling the engine 1 for performing the SPCCI combustion,
the close timing IVC of the intake valve 21 is set based on the
relation between the geometric compression ratio .epsilon. and the
close timing IVC of the intake valve 21.
The engine 1 switches between "Layer 2" where the SPCCI combustion
is performed with the A/F of the mixture gas being made at the
stoichiometric air fuel ratio or richer than the stoichiometric air
fuel ratio and the G/F being made leaner than the stoichiometric
air fuel ratio, and "Layer 3" where the SPCCI combustion is
performed with the A/F of the mixture gas being made leaner than
the stoichiometric air fuel ratio. The relation between the
geometric compression ratio .epsilon. and the close timing IVC of
the intake valve 21 in Layer 2 differs from the relation between
the geometric compression ratio .epsilon. and the close timing IVC
of the intake valve 21 in Layer 3.
(1-1) Relation Between Geometric Compression Ratio and Close Timing
of Intake Valve in Layer
FIG. 15 illustrates the characteristic of the SPCCI combustion. For
example, FIG. 15 illustrates the occurring range of the SPCCI
combustion against the EGR ratio (horizontal axis) in a case where
the engine 1 carries out the SPCCI combustion of the mixture gas of
which the A/F is the stoichiometric air fuel ratio and the G/F is
leaner than the stoichiometric air fuel ratio, similar to Layer 2.
The vertical axis in this graph is a crank angle corresponding to a
combustion center of gravity, where the combustion center of
gravity advances as it goes upward in this graph.
The occurring range of the SPCCI combustion is illustrated by a
hatched range in this graph. The occurring range of the SPCCI
combustion is located between a line of "advancing limit" and a
line of "retarding limit." If the combustion center of gravity
advances beyond the "advancing limit" line, combustion becomes
abnormal, which means that the SPCCI combustion does not occur.
Similarly, if the combustion center of gravity retards beyond the
"retarding limit" line, self-ignition does not occur, which means
that the SPCCI combustion does not occur.
The one-dot chain line in this graph illustrates the combustion
center of gravity of combustion corresponding to a MBT (Minimum
advance for Best Torque). Here, the combustion center of gravity of
combustion corresponding to the MBT is simply referred to as "MBT."
The MBT is advanced as the EGR ratio increases.
In terms of the improvement in fuel efficiency, it is desirable to
bring the combustion center of gravity of the SPCCI combustion
closer to the MBT. As the EGR ratio increases, the occurring range
of the SPCCI combustion is advanced, but the interval between the
"advancing limit" line and the "retarding limit" line becomes
narrower, which narrows the occurring range of the SPCCI
combustion.
When the engine load is low (when "light load"), the occurring
range of the SPCCI combustion is on the advance side. Therefore, as
illustrated by a both-ends arrow in FIG. 15, the SPCCI combustion
corresponding to the MBT can be achieved by adjusting the EGR ratio
within a certain width, and adjusting the combustion center of
gravity to the advance side or the retard side.
When the engine load increases, the amount of air introduced into
the combustion chamber 17 must be increased corresponding to the
increase in the fuel supply amount. When the EGR ratio is increased
corresponding to the increase in the air amount, a large amount of
the EGR gas must be introduced into the combustion chamber 17.
However, because of the limit of the supercharging capability of
the supercharger 44, it is difficult to introduce both a large
amount of air and a large amount of the EGR gas into the combustion
chamber 17. Therefore, when the engine load increases, the
occurring range of the SPCCI combustion becomes on the retard side.
If bringing the combustion center of gravity of the SPCCI
combustion closest to the MBT when the engine load is high, the
SPCCI combustion must be performed at a point Y in FIG. 15. The
point Y corresponds to the operating state of the engine 1 with the
maximum load where the SPCCI combustion of the mixture gas with the
A/F being at the stoichiometric air fuel ratio is possible. When
the engine 1 operates with the maximum load, it is difficult to
achieve the SPCCI combustion corresponding to the MBT by adjusting
the EGR ratio or the combustion center of gravity.
The engine 1 operates over a wide operating range from the low load
to the high load in Layer 2. In Layer 2, the state where the engine
1 operates with the maximum load corresponds to the operating state
at the limit where the SPCCI combustion can occur. In Layer 2, upon
determining the relation between .epsilon. and IVC so that the
temperature of the combustion chamber 17 at the start timing
(.theta..sub.CI) in the CI combustion becomes the given
temperature, it is necessary to set the relation based on the
temperature inside the combustion chamber 17 when the engine 1
operates with the maximum load.
In order to obtain the temperature inside the combustion chamber 17
when the engine 1 operates with the maximum load, the present
inventors performed the SPCCI combustion in an actual engine 1, and
used measurements acquired from the engine 1. For example, the
present inventors measured various parameters when the engine 1
operates with the maximum load in Layer 2, and estimated an actual
temperature inside the combustion chamber 17 at .theta..sub.CI
based on the measured parameter. The present inventors used an
average value of a plurality of estimated temperatures as a
reference temperature Tth1. If the temperature of the combustion
chamber 17 at .theta..sub.CI is the reference temperature Tth1, the
SPCCI combustion can be achieved in Layer 2.
Here, as described above, .theta..sub.CI an be adjusted by
adjusting the ignition timing in the SPCCI combustion. However,
while the engine 1 operates with the maximum load in Layer 2, it
becomes impossible to adjust .theta..sub.CI even if the ignition
timing is adjusted when the temperature inside the combustion
chamber 17 at .theta..sub.CI exceeds the reference temperature
Tth1. On the other hand, if the temperature inside the combustion
chamber 17 at .theta..sub.CI is the reference temperature Tth1 or
lower, the ignition timing is adjusted (e.g., the ignition timing
is advanced) to raise the temperature inside the combustion chamber
17 at .theta..sub.CI to the reference temperature Tth1, thereby
achieving the SPCCI combustion.
Therefore, in order to achieve the SPCCI combustion in Layer 2, the
relation between .epsilon. and IVC is required so that the
temperature of the combustion chamber 17 at .theta..sub.CI does not
exceed the reference temperature Tth1.
Thus, as conceptually illustrated in FIG. 16, the present inventors
performed an estimation of the temperature of the combustion
chamber 17 at .theta..sub.CI by using a model of the engine 1,
while changing the values of the geometric compression ratio
.epsilon. and the close timing IVC of the intake valve 21 in a
matrix comprised of the two parameters of .epsilon. and IVC. A
combination of and IVC where the temperature of the combustion
chamber 17 becomes the reference temperature Tth1 or lower can
achieve the SPCCI combustion in Layer 2. As illustrated in FIG. 16,
when the geometric compression ratio .epsilon. is high and the
close timing IVC of the intake valve 21 approaches an intake bottom
dead center, the temperature of the combustion chamber 17 during
the CI combustion exceeds the reference temperature Tth1.
Note that FIG. 16 illustrates the matrix where IVC is set after the
intake bottom dead center. Although illustration is omitted, the
present inventors also acquired a combination of and IVC where the
temperature becomes the reference temperature Tth1 or lower, by
performing an estimation of the temperature of the combustion
chamber 17 at .theta..sub.CI, similarly for the matrix where IVC is
set before the intake bottom dead center.
A graph 1701 of the upper figure in FIG. 17 illustrates
approximations (I) and (II) calculated based on the combination of
.epsilon. and IVC. The horizontal axis of the graph 1701 is the
geometric compression ratio .epsilon., and the vertical axis is the
close timing IVC (deg.aBDC) of the intake valve 21. Although
illustration is omitted, the present inventors plotted on the graph
1701 the combination of .epsilon. and IVC where the temperature
becomes the reference temperature Tth1 or lower, and determined the
approximations (I) and (II) based on the plots.
The graph 1701 corresponds to a case where the engine speed is
2,000 rpm. The approximations (I) and (II) are as follows. IVC
=-0.4288.epsilon..sup.2+31.518.epsilon.-379.88 Approximation (I)
IVC=-1.9163.epsilon..sup.2-89.935.epsilon.+974.94 Approximation
(II)
In the graph 1701, the combination of .epsilon. and IVC on the left
side of the approximations (I) and (II) and .epsilon.=17, the
temperature of the combustion chamber 17 during the CI combustion
becomes the reference temperature Tth1 or lower. In this
combination, it is possible to carry out the SPCCI combustion of
the mixture gas with the A/F being the stoichiometric air fuel
ratio and the G/F being leaner than the stoichiometric air fuel
ratio.
The relation between .epsilon. and IVC described above is a
relation based on the maximum temperature of the combustion chamber
17 in Layer 2.
On the other hand, in Layer 2, also while the engine 1 operates
with the light load, the relation between .epsilon. and IVC must be
defined so that the temperature of the combustion chamber 17
becomes the given temperature.
The temperature of the combustion chamber 17 when the SPCCI
combustion is performed is a result of two pressure buildups of a
pressure buildup by the compression work of the piston 3 in a
compression stroke, and the pressure buildup caused by the heat
generation of the SI combustion. The compression work of the piston
3 is determined by the effective compression ratio. If the
effective compression ratio is too low, the pressure buildup by the
compression work of the piston 3 decreases. Unless the pressure
buildup, which is caused by the heat generation of the SI
combustion after the flame propagation in the SPCCI combustion
progresses, increases considerably, the in-cylinder temperature
cannot be raised to an ignition temperature. As a result, since the
amount of the mixture gas which is ignited by the compressed
self-ignition is little, and most of the mixture gas burn by the
flame propagation, the combustion period becomes longer and fuel
efficiency falls. In order to stabilize the CI combustion in the
SPCCI combustion and maximize fuel efficiency, it is necessary to
maintain the effective compression ratio above a certain value.
Therefore, the relation between .epsilon. and IVC must be
determined accordingly.
Similarly, the present inventors measured various parameters when
the actual engine 1 operates with the light load, and estimated an
actual temperature inside the combustion chamber 17 at
.theta..sub.CI based on the measured parameters. The present
inventors used an average value of a plurality of estimated
temperatures as a reference temperature Tth2.
If the temperature inside the combustion chamber 17 at
.theta..sub.CI is the reference temperature Tth2 or higher while
the engine 1 operates with the light load, the SPCCI combustion can
be achieved by delaying the ignition timing. However, since the
temperature of the combustion chamber 17 is too low if the
temperature at .theta..sub.CI is lower than the reference
temperature Tth2, the SPCCI combustion cannot be achieved even if
the ignition timing is advanced.
Thus, in order to achieve the SPCCI combustion in Layer 2, a
relation between .epsilon. and IVC is required so that the
temperature of the combustion chamber 17 at .theta..sub.CI becomes
the reference temperature Tth2 or higher.
The present inventors performed an estimation of the temperature of
the combustion chamber 17 during the CI combustion by using a model
of the engine 1, while changing the values of the geometric
compression ratio .epsilon. and the close timing IVC of the intake
valve 21 in a matrix comprised of two parameters of .epsilon. and
IVC, similarly to the matrix illustrated in FIG. 16. In this
matrix, the combination of .epsilon. and IVC where the temperature
of the combustion chamber 17 becomes the reference temperature Tth2
or higher can achieve the SPCCI combustion in Layer 2.
In the graph 1701 of FIG. 17, approximations (III) and (IV)
calculated based on the combination of .epsilon. and IVC where the
temperature becomes the reference temperature Tth2 or higher are
also illustrated. The approximations (III) and (IV) are as follows.
IVC=-0.4234.epsilon..sup.2+22.926.epsilon.-167.84 Approximation
(III) IVC=0.4234.epsilon..sup.2-22.926.epsilon.+207.84
Approximation (IV)
In the graph 1701, the combination of .epsilon. and IVC on the
right side of the approximations (III) and (IV), the temperature of
the combustion chamber 17 during the CI combustion becomes the
reference temperature Tth2 or higher. In this combination, the
SPCCI combustion of the mixture gas with the A/F being the
stoichiometric air fuel ratio and the G/F being leaner than the
stoichiometric air fuel ratio is possible.
As seen in FIG. 17, the relation between .epsilon. and IVC is
substantially vertically symmetrical with respect to IVC=about 20
deg.aBDC. IVC=20 deg.aBDC corresponds to a close timing (i.e., the
best IVC) at which the amount of gas introduced into the combustion
chamber 17 becomes the maximum when the engine speed is 2,000 rpm.
Moreover, IVC=120 deg.aBDC is a retarding limit of the close timing
IVC of the intake valve 21, and IVC=-80 deg.aBDC is an advancing
limit of the close timing IVC of the intake valve 21.
The combination of .epsilon. and IVC within a range surrounded by
the approximations (I), (II), (III), and (IV) in FIG. 17 is a
combination which can put in practical use the engine 1 for
performing the SPCCI combustion in Layer 2. In other words, the
combination of .epsilon. and IVC outside this range cannot put in
practical use the engine 1 for performing the SPCCI combustion in
Layer 2.
The ECU 10 has to determine IVC within the .epsilon.-IVC valid
range hatched in FIG. 17, upon determining the close timing IVC of
the intake valve 21 when the engine 1 operates in Layer 2.
For example, if the geometric compression ratio .epsilon. is set as
10.ltoreq..epsilon.<17, the ECU 10 determines the close timing
IVC (deg.aBDC) so that the following expression is satisfied.
0.4234.epsilon..sup.2-22.926.epsilon.+207.84.ltoreq.IVC.ltoreq.-0.4234.ep-
silon..sup.2+22.926.epsilon.-167.84 (1)
Moreover, if the geometric compression ratio .epsilon. is set as
17.ltoreq..epsilon.<20, the ECU 10 determines the close timing
IVC (deg.aBDC) so that the following expression is satisfied.
-0.4288.epsilon..sup.2+31.518.epsilon.-379.88.ltoreq.IVC.ltoreq.-0.4234.e-
psilon..sup.2+22.926.epsilon.-167.84 (2) or
0.4234.epsilon..sup.2-22.926.epsilon.+207.84.ltoreq.IVC.ltoreq.1.9163.eps-
ilon..sup.2-89.935.epsilon.+974.94 (3)
Further, if the geometric compression ratio .epsilon. is set as
20.ltoreq..epsilon..ltoreq.30, the ECU 10 determines the close
timing IVC (deg.aBDC) so that the following expression is
satisfied.
-0.4288.epsilon..sup.2+31.518.epsilon.-379.88.ltoreq.IVC.ltoreq.120
(4) or
-80.ltoreq.IVC.ltoreq.1.9163.epsilon..sup.2-89.935.epsilon.+974.94
(5)
By setting the close timing IVC of the intake valve 21 based on the
relational expressions (1) to (5), the SPCCI combustion of the
mixture gas with the A/F being the stoichiometric air fuel ratio or
richer than the stoichiometric air fuel ratio and the G/F being
leaner than the stoichiometric air fuel ratio is achieved. Note
that the close timing IVC is set for each operating state which is
determined by the load and the engine speed in Layer 2. The example
illustrated by the solid line in FIG. 17 is the .epsilon.-IVC valid
range when the engine speed is 2,000 rpm, as described above. If
the engine speed changes, the .epsilon.-IVC valid range also
changes. As the engine speed increases, the best IVC is
retarded.
For example, when the engine speed is 3,000 rpm, the best IVC is
about 22 deg.aBDC. As illustrated by broken lines in FIG. 17, the
.epsilon.-IVC valid range when the engine speed is 3,000 rpm can be
obtained by parallelly translating the .epsilon.-IVC valid range
when the engine speed is 2,000 rpm to the retard side by about 2
degrees.
Therefore, if the geometric compression ratio .epsilon. is set as
10.ltoreq..epsilon.<17, when the engine speed is 3,000 rpm, the
ECU 10 determines the close timing IVC (deg.aBDC) so that the
following expression is satisfied.
0.4234.epsilon..sup.2-22.926.epsilon.+209.84.ltoreq.IVC.ltoreq.-0.4234.ep-
silon..sup.2+22.926.epsilon.-165.84 (1.sup.-1)
Moreover, if the geometric compression ratio .epsilon. is set as
17.ltoreq..epsilon.<20, when the engine speed is 3,000 rpm, the
ECU 10 determines the close timing IVC (deg.aBDC) so that the
following expression is satisfied.
-0.4288.epsilon..sup.2+31.518.epsilon.-377.88.ltoreq.IVC.ltoreq.-0.4234.e-
psilon..sup.2+22.926.epsilon.-165.84 (2.sup.-1) or
0.4234.epsilon..sup.2-22.926.epsilon.+209.84.ltoreq.IVC.ltoreq.1.9163.eps-
ilon..sup.2-89.935.epsilon.+976.94 (3.sup.-1)
Further, if the geometric compression ratio .epsilon. is set as
20.ltoreq..epsilon..ltoreq.<30, the ECU 10 determines the close
timing IVC (deg.aBDC) so that the following expression is
satisfied.
-0.4288.epsilon..sup.2+31.518.epsilon.-377.88.ltoreq.IVC.ltoreq.120
(4.sup.-1) or
-80.ltoreq.IVC.ltoreq.1.9163.epsilon..sup.2-87.935.epsilon.+976.94
(5.sup.-1)
Moreover, when the engine speed is 4,000 rpm, the best IVC is about
28 deg.aBDC. As illustrated by one-dot chain lines in FIG. 17, the
.epsilon.-IVC valid range when the engine speed is 4,000 rpm is
obtained by parallelly translating the .epsilon.-IVC valid range
when the engine speed is 2,000 rpm to the retard side by about 8
degrees.
Therefore, if the geometric compression ratio .epsilon. is set as
10.ltoreq..epsilon.<17, when the engine speed is 4,000 rpm, the
ECU 10 determines the close timing IVC (deg.aBDC) so that the
following expression is satisfied.
0.4234.epsilon..sup.2-22.926.epsilon.+215.84.ltoreq.IVC.ltoreq.-0.4234.ep-
silon..sup.2+22.926.epsilon.-159.84 (1.sup.-2)
Moreover, if the geometric compression ratio .epsilon. is set as
17.ltoreq..epsilon.<20, when the engine speed is 4,000 rpm, the
ECU 10 determines the close timing IVC (deg.aBDC) so that the
following expression is satisfied.
-0.4288.epsilon..sup.2+31.518.epsilon.-371.88.ltoreq.IVC.ltoreq.-0.4234.e-
psilon..sup.2+22.926.epsilon.-159.84 (2.sup.-2) or
0.4234.epsilon..sup.2-22.926.epsilon.+215.84.ltoreq.IVC.ltoreq.1.9163.eps-
ilon..sup.2-89.935.epsilon.+982.94 (3.sup.-2)
Further, if the geometric compression ratio .epsilon. is set as
20.ltoreq..epsilon..ltoreq.30, the ECU 10 determines the close
timing IVC (deg.aBDC) so that the following expression is
satisfied.
-0.4288.epsilon..sup.2+31.518.epsilon.-371.88.ltoreq.IVC.ltoreq.120
(4.sup.-2) or
-80.ltoreq.IVC.ltoreq.1.9163.epsilon..sup.2-89.935.epsilon.+982.94
(5.sup.-2)
If the correction term C according to the engine speed NE (rpm) of
the engine 1 is set as the following,
C=3.3.times.10.sup.-NE.sup.3-1.0.times.10.sup.-6NE.sup.3+7.0.times.10.sup-
.-4NE the relational expression of .epsilon. and IVC in Layer 2 can
be expressed as follows. If the geometric compression ratio
.epsilon. is 10.ltoreq..epsilon.<17,
0.4234.epsilon..sup.2-22.926.epsilon.+207.84+C.ltoreq.IVC.ltoreq.-0.4234.-
epsilon..sup.2+22.926.epsilon.-167.84+C (1.sup.-3) If the geometric
compression ratio .epsilon. is 17.ltoreq..epsilon.<20,
0.4288.epsilon..sup.2+31.518.epsilon.-379.88+C.ltoreq.IVC.ltoreq.-0.4234.-
epsilon..sup.2+22.926.epsilon.-167.84+C (2.sup.-3) or
0.4234.epsilon..sup.2-22.926.epsilon.+207.84+C.ltoreq.IVC.ltoreq.1.9163.e-
psilon..sup.2-89.935.epsilon.+974.94+C (3.sup.-3) If the geometric
compression ratio .epsilon. is 20.ltoreq..epsilon..ltoreq.30,
-0.4288.epsilon..sup.2+31.518.epsilon.-379.88+C.ltoreq.IVC.ltoreq.120
(4.sup.-3) or
-80.ltoreq.IVC.ltoreq.1.9163.epsilon..sup.2-89.935.epsilon.+974.94+C
(5.sup.-3) The ECU 10 determines the close timing IVC based on the
.epsilon.-IVC valid range determined for every engine speed of the
engine 1. As a result, the ECU 10 can set the valve timing of the
intake valve 21 in Layer 2 as illustrated in FIG. 11. (1-2) Change
in .epsilon.-IVC Valid Range by Difference of Octane Number
The graph 1701 in FIG. 17 is a relation between .epsilon. and IVC
when the fuel is high octane fuel (octane number is about 96). A
graph 1702 illustrated in the lower figure is a relation between
.epsilon. and IVC when the fuel is low octane fuel (octane number
is about 91). According to the examination of the present
inventors, when the fuel was the low octane fuel, it was found that
the .epsilon.-IVC valid range shifts by 1.3 compression ratios
toward the lower compression ratio from the .epsilon.-IVC valid
range of the high octane fuel.
Accordingly, upon determining the close timing IVC in the engine 1
of the low octane fuel, if the geometric compression ratio
.epsilon. is set as 10.ltoreq..epsilon.<15.7, when the engine
speed is 2,000 rpm, the ECU 10 determines the close timing IVC
(deg.aBDC) so that the following expression is satisfied.
0.4234.epsilon..sup.2-21.826.epsilon.+178.75.ltoreq.IVC.ltoreq.-0.4234.ep-
silon..sup.2+21.826.epsilon.-138.75 (6)
Moreover, if the geometric compression ratio .epsilon. is set as
15.7.ltoreq..epsilon.<18.7 in the engine 1 of the low octane
fuel, when the engine speed is 2,000 rpm, the ECU 10 determines the
close timing IVC (deg.aBDC) so that the following expression is
satisfied.
-0.5603.epsilon..sup.2+34.859.epsilon.-377.22.ltoreq.IVC.ltoreq.-0.4234.e-
psilon..sup.2+21.826.epsilon.-138.75 (7) or
0.4234.epsilon..sup.2-21.826.epsilon.+178.75.ltoreq.IVC.ltoreq.1.9211.eps-
ilon..sup.2-85.076.epsilon.+862.01 (8)
Further, if the geometric compression ratio .epsilon. is set as
18.7.ltoreq..epsilon..ltoreq.30 in the engine 1 of the low octane
fuel, when the engine speed is 2,000 rpm, the ECU 10 determines the
close timing IVC (deg.aBDC) so that the following expression is
satisfied.
0.5603.epsilon..sup.2+34.859.epsilon.-377.22.ltoreq.IVC.ltoreq.120
(9) or
-80.ltoreq.IVC.ltoreq.1.9211.epsilon..sup.2-85.076.epsilon.+862.01
(10)
The cross-hatched range in the graph 1702 of FIG. 17 is an
overlapping range of the .epsilon.-IVC valid range of the high
octane fuel and the .epsilon.-IVC valid range of the low octane
fuel. The ECU 10 can set the control logic which suits both the
engine 1 using the high octane fuel and the engine 1 using the low
octane fuel, if IVC is determined within the overlapping range of
the two valid ranges. Even if the octane number of the fuel differs
for every destination of this product, the ECU 10 can collectively
use the engine control logic.
Note that although illustration is omitted, the .epsilon.-IVC valid
range is parallelly translated to the retard side also in the
engine 1 of the low octane fuel, when the engine speed increases.
If the geometric compression ratio .epsilon. is set as
10.ltoreq..epsilon.<15.7 in the engine 1 of the low octane fuel,
when the engine speed is 3,000 rpm, the ECU 10 determines the close
timing IVC (deg.aBDC) so that the following expression is
satisfied.
0.4234.epsilon..sup.2-21.826.epsilon.+180.75.ltoreq.IVC.ltoreq.-0.4234.ep-
silon..sup.2+21.826.epsilon.-136.75 (6.sup.-1)
Moreover, if the geometric compression ratio .epsilon. is set as
15.7.ltoreq..epsilon.<18.7 in the engine 1 of the low octane
fuel, when the engine speed is 3,000 rpm, the ECU 10 determines the
close timing IVC (deg.aBDC) so that the following expression is
satisfied.
-0.5603.epsilon..sup.2+34.859.epsilon.-375.22.ltoreq.IVC.ltoreq.-0.4234.e-
psilon..sup.2+21.826.epsilon.-136.75 (7.sup.-1) or
0.4234.epsilon..sup.2-21.826.epsilon.+180.75.ltoreq.IVC.ltoreq.1.9211.eps-
ilon..sup.2-85.076.epsilon.+864.01 (8.sup.-1)
Further, if the geometric compression ratio .epsilon. is set as
18.7.ltoreq..epsilon.<30 in the engine 1 of the low octane fuel,
when the engine speed is 3,000 rpm, the ECU 10 determines the close
timing IVC (deg.aBDC) so that the following expression is
satisfied.
-0.5603.epsilon..sup.2+34.859.epsilon.-375.22.ltoreq.IVC.ltoreq.120
(9.sup.-1) or
-80.ltoreq.IVC.ltoreq.1.9211.epsilon..sup.2-85.076.epsilon.+864.01
(10.sup.-1)
Moreover, if the geometric compression ratio .epsilon. is set as
10.ltoreq..epsilon.<15.7 in the engine 1 of the low octane fuel,
when the engine speed is 4,000 rpm, the ECU 10 determines the close
timing IVC (deg.aBDC) so that the following expression is
satisfied.
0.4234.epsilon..sup.2-21.826.epsilon.+186.75.ltoreq.IVC.ltoreq.-0.4234.ep-
silon..sup.2+21.826.epsilon.-130.75 (6.sup.-2)
Moreover, if the geometric compression ratio .epsilon. is set as
15.7.ltoreq..epsilon.<18.7 in the engine 1 of the low octane
fuel, when the engine speed is 4,000 rpm, the ECU 10 determines the
close timing IVC (deg.aBDC) so that the following expression is
satisfied.
-0.5603.epsilon..sup.2+34.859.epsilon.-369.22.ltoreq.IVC.ltoreq.-0.4234.e-
psilon..sup.2+21.826.epsilon.-130.75 (7.sup.-2) or
0.4234.epsilon..sup.2-21.826.epsilon.+186.75.ltoreq.IVC.ltoreq.1.9211.eps-
ilon..sup.2-85.076.epsilon.+870.01 (8.sup.-2)
Further, if the geometric compression ratio .epsilon. is set as
18.7.ltoreq..epsilon.<30 in the engine 1 of the low octane fuel,
when the engine speed is 4,000 rpm, the ECU 10 determines the close
timing IVC (deg.aBDC) so that the following expression is
satisfied.
0.5603.epsilon..sup.2+34.859.epsilon.-369.22.ltoreq.IVC.ltoreq.120
(9.sup.-2) or
-80.ltoreq.IVC.ltoreq.1.9211.epsilon..sup.2-77.076.epsilon.+870.01
(10.sup.-2)
If the correction term C according to the engine speed NE (rpm) of
the engine 1 is used similar to the above, the relational
expression of .epsilon. and IVC in Layer 2 in the engine 1 of the
low octane fuel can be expressed as follows.
If the geometric compression ratio .epsilon. is
10.ltoreq..epsilon..ltoreq.15.7,
0.4234.epsilon..sup.2-21.826.epsilon.+178.75+C.ltoreq.IVC.ltoreq.-0.4234.-
epsilon..sup.2+21.826.epsilon.-138.75+C (6.sup.-3) If the geometric
compression ratio .epsilon. is 15.7.ltoreq..epsilon.<18.7,
-0.5603.epsilon..sup.2+34.859.epsilon.-377.22+C.ltoreq.IVC.ltoreq.-0.4234-
.epsilon..sup.2+21.826.epsilon.-138.75+C (7.sup.-3) or
0.4234.epsilon..sup.2-21.826.epsilon.+178.75+C.ltoreq.IVC.ltoreq.1.9211.e-
psilon..sup.2-85.076.epsilon.+862.01+C (8.sup.-3) If the geometric
compression ratio .epsilon. is 18.7.ltoreq..epsilon..ltoreq.30,
-0.5603.epsilon..sup.2+34.859.epsilon.-377.22+C.ltoreq.IVC.ltoreq.120
(9.sup.-3) or
-80.ltoreq.IVC.ltoreq.1.9211.epsilon..sup.2-85.076.epsilon.+862.01+C
(10.sup.-3) (1-3) Relation Between Geometric Compression Ratio and
Close Timing of Intake Valve in Layer 3
FIG. 18 illustrates the characteristic of the SPCCI combustion when
the engine 1 carries out the SPCCI combustion of the mixture gas
where the A/F is leaner than the stoichiometric air fuel ratio,
similar to Layer 3. FIG. 18 illustrates a range where the SPCCI
combustion is stable with respect to the G/F (horizontal axis). The
vertical axis in this graph is a crank angle corresponding to a
combustion center of gravity, and the combustion center of gravity
is advanced as it goes upward in this graph.
The range where the SPCCI combustion is stabilized is a range
surrounded by a curve in this graph. When the engine load is low,
the range where the SPCCI combustion is stabilized is located at
upper left in FIG. 18. When the engine load increases, the range
where the SPCCI combustion is stabilized moves downwardly in FIG.
18.
FIG. 18 also illustrates a range where discharge of raw NO.sub.x
can be reduced. The range where the discharge of raw NO.sub.x can
be reduced is located at lower right in FIG. 18. This range has a
triangular shape in FIG. 18. If the A/F is leaner than the
stoichiometric air fuel ratio, raw NO.sub.x cannot be purified by
the three-way catalyst. In Layer 3, the engine 1 must satisfy both
securing of the stability of the SPCCI combustion and reduction of
the discharge of raw NO.sub.x.
As seen in this graph, if the engine load is high, an overlapping
area of the range where the combustion stability is secured and the
range where the discharge of raw NO.sub.x can be reduced increases.
On the other hand, if the engine load is low, the overlapping area
of the range where the combustion stability is secured and the
range where the discharge of raw NO.sub.x can be reduced
decreases.
Regarding Layer 3, the state where the engine 1 operates with the
light load corresponds to a limit of the operating state where the
SPCCI combustion can occur. Regarding Layer 3, upon determining the
relation between .epsilon. and IVC so that the temperature of the
combustion chamber 17 at the start timing (.theta..sub.CI) of the
CI combustion becomes the given temperature, it is necessary to be
set based on the temperature inside the combustion chamber 17 when
the engine 1 operates with the light load.
Similar to the above, using the actual engine 1, the present
inventors measured various parameters when operating the engine 1
with the light load in Layer 3, and estimated an actual temperature
inside the combustion chamber 17 at .theta..sub.CI based on the
measured parameters. Then, an average value of a plurality of
estimated temperatures was determined as a reference temperature
Tth3. If the temperature of the combustion chamber 17 at
.theta..sub.CI is the reference temperature Tth3, the SPCCI
combustion can be achieved in Layer 3. This reference temperature
Tth3 corresponds to the minimum temperature. If the temperature
inside the combustion chamber 17 at .theta..sub.CI is the reference
temperature Tth3 or higher while the engine 1 operates with the
light load, the SPCCI combustion can be achieved by delaying the
ignition timing. However, if the temperature inside the combustion
chamber 17 at .theta..sub.CI is lower than the reference
temperature Tth3, SPCCI combustion cannot be achieved even if the
ignition timing is advanced.
Therefore, in order to achieve the stable SPCCI combustion in Layer
3, the relation between .epsilon. and IVC is required so that the
temperature inside the combustion chamber 17 at .theta..sub.CI
becomes the reference temperature Tth3 or higher.
Thus, as conceptually illustrated in FIG. 19, the present inventors
estimated the temperature of the combustion chamber 17 during the
CI combustion by using the model of the engine 1, while changing
the values of IVC and O/L in a matrix of the close timing IVC of
the intake valve 21 and the overlap period O/L of the intake valve
21 and the exhaust valve 22, for every geometric compression ratio
.epsilon. (.epsilon.1, .epsilon.2 . . . ) (a reference numeral
1901). As hatched in the matrix of the reference numeral 1901, a
combination of IVC and O/L where the temperature becomes the
reference temperature Tth3 or higher can achieve the SPCCI
combustion in Layer 3.
Moreover, in order to reduce the discharge of raw NO.sub.x, the G/F
of the mixture gas must be made a given value or higher. As
illustrated by a reference numeral 1902 in FIG. 19, the present
inventors estimated the G/F by using the model of the engine 1,
while changing the values of the close timing IVC and the overlap
period O/L in the matrix comprised of the two parameters of IVC and
O/L. As oblique lines are drawn in the matrix of the reference
numeral 1902, the combination of IVC and O/L where the G/F becomes
the given value or higher can reduce the discharge of raw
NO.sub.x.
Then, the present inventors determined the relation between
.epsilon. and IVC which can achieve both the stability of the SPCCI
combustion and the reduction of raw NO.sub.x discharge, by
overlapping the combination of IVC and O/L where the temperature
becomes the reference temperature Tth3 or higher which is
illustrated by the reference numeral 1901, and the combination of
IVC and O/L where the G/F becomes the given value or higher which
is illustrated by the reference numeral 1902. That is, in the
matrix of a reference numeral 1903, a cross-hatched range is the
combination of .epsilon. and IVC which can achieve both the
stability of the SPCCI combustion and the reduction of NO.sub.x
discharge.
Note that although illustration is omitted, the present inventors
acquired a combination of IVC and O/L where the temperature becomes
the reference temperature Tth3 or higher and a combination of IVC
and O/L where the G/F becomes the given value or higher by
estimating the temperature of the combustion chamber 17 and the G/F
during the CI combustion, similarly for the matrix where the close
timing of the intake valve 21 is set before the intake bottom dead
center.
FIG. 20 illustrates approximations (V) and (VI) calculated based on
the combinations of .epsilon. and IVC. The horizontal axis in FIG.
20 is the geometric compression ratio .epsilon., and the vertical
axis is the close timing IVC (deg. aBDC) of the intake valve
21.
An upper figure 2001 of FIG. 20 corresponds to a case when the
engine speed is 2,000 rpm. The approximations (V) and (VI) are as
follows. IVC=-0.9949.epsilon..sup.2+41.736.epsilon.-361.16
Approximation (V) IVC=0.9949.epsilon..sup.2-41.736.epsilon.+401.16
Approximation (VI)
In FIG. 20, the combinations of .epsilon. and IVC on the right side
of the approximations (V) and (VI) have the temperature of the
combustion chamber 17 during CI combustion becoming the reference
temperature Tth3 or higher, thereby achieving the SPCCI combustion
of the mixture gas with the A/F being leaner than the
stoichiometric air fuel ratio.
In Layer 3, the relation between .epsilon. and IVC described above
is a relation based on the minimum temperature of the combustion
chamber 17 which can achieve the SPCCI combustion when the engine 1
operates with the light load.
On the other hand, if the temperature inside the combustion chamber
17 is too high, the CI combustion begins before the start of the SI
combustion and the SPCCI combustion cannot be performed, regardless
of Layer 2 or Layer 3.
Here, the concept of the SPCCI combustion is such that, as
described above, when the ignition plug 25 ignites the mixture gas,
the mixture gas around the ignition plug 25 starts the SI
combustion, and, after that, the surrounding mixture gas carries
out the CI combustion. From examinations by experiments, etc. which
the present inventors conducted until now, it was found that the
self-ignition of the mixture gas occurred when the surrounding
temperature of the mixture gas which self-ignites exceeds a given
reference temperature Tth4, and this reference temperature Tth4 was
not necessarily in agreement with a mean temperature of the entire
combustion chamber 17. From this knowledge, if the mean temperature
inside the combustion chamber 17 at a compression top dead center
reaches the reference temperature Tth4, it is thought that the CI
combustion will begin before the SI combustion begins, and in this
case, the SPCCI combustion cannot be performed.
Thus, the present inventors estimated the temperature of the
combustion chamber 17 at the compression top dead center by using
the model of the engine 1, while changing the values of the close
timing IVC and the overlap period O/L in the matrix of the close
timing IVC of the intake valve 21, and the overlap period O/L of
the intake valve 21 and the exhaust valve 22, for every geometric
compression ratio .epsilon. (.epsilon.1, .epsilon.2 . . . ),
similar to the matrix of the reference numeral 1901 of FIG. 19. The
combination of IVC and O/L where the temperature inside the
combustion chamber 17 at the compression top dead center exceeds
the reference temperature Tth4 cannot achieve the SPCCI combustion,
but the combination of IVC and O/L being the reference temperature
Tth4 or lower can achieve the SPCCI combustion.
FIG. 20 illustrates approximations (VII) and (VIII) calculated
based on the combination of .epsilon. and IVC where the temperature
inside the combustion chamber 17 at the compression top dead center
becomes the reference temperature Tth4 or lower. The approximations
(VII) and (VIII) are as follows.
IVC=-4.7481.epsilon..sup.2+266.75.epsilon.-3671.2 Approximation
(VII) and IVC=4.7481.epsilon..sup.2-266.75.epsilon.+3711.2
Approximation (VIII)
In FIG. 20, the combination of .epsilon. and IVC on the left side
the approximations (VII) and (VIII) can avoid that the CI
combustion begins before the SI combustion, and achieves the SPCCI
combustion.
As seen in FIG. 20, also in Layer 3, the relation between .epsilon.
and IVC is substantially vertically symmetrical with respect to
IVC=20 deg.aBDC. Moreover, IVC=75 deg.aBDC is a retarding limit of
the close timing of the intake valve 21 set in consideration of the
amount of gas introduced into the combustion chamber 17 when the
engine 1 operates in Layer 3. Similarly, IVC=-35 deg.aBDC is an
advancing limit of the close timing of the intake valve 21 set in
consideration of the amount of gas introduced into the combustion
chamber 17.
When determining the close timing IVC of the intake valve 21 in the
case where the engine 1 operates in Layer 3, the ECU 10 must
determine IVC within the .epsilon.-IVC valid range surrounded by
the approximations (V), (VI), (VII), and (VIII) in FIG. 20 (the
hatched range in FIG. 20).
For example, if the geometric compression ratio .epsilon. is set as
10.ltoreq..epsilon.<20, when the engine speed is 2,000 rpm, the
ECU 10 determines the close timing IVC (deg.aBDC) so that the
following expression is satisfied.
0.9949.epsilon..sup.2-41.736.epsilon.+401.16.ltoreq.IVC.ltoreq.-0.9949.ep-
silon..sup.2+41.736.epsilon.-361.16 (11)
Moreover, if the geometric compression ratio .epsilon. is set as
20.ltoreq..epsilon.<25, when the engine speed is 2,000 rpm, the
ECU 10 determines the close timing IVC (deg.aBDC) so that the
following expression is satisfied. -35.ltoreq.IVC.ltoreq.75
(12)
Further, if the geometric compression ratio .epsilon. is set as
25.ltoreq..epsilon..ltoreq.30, when the engine speed is 2,000 rpm,
the ECU 10 determines the close timing IVC (deg.aBDC) so that the
following expression is satisfied.
-4.7481.epsilon..sup.2+266.75.epsilon.-3671.2.ltoreq.IVC.ltoreq.75
(13) or
-35.ltoreq.IVC.ltoreq.4.7481.epsilon..sup.2-266.75.epsilon.+3711.2
(14)
By setting the close timing IVC of the intake valve 21 based on the
relational expressions (11) to (14), the SPCCI combustion is
achieved of the mixture gas with a A/F leaner than the
stoichiometric air fuel ratio. Note that the close timing IVC is
set for each operating state which is determined by the load and
the engine speed in Layer 3.
The example illustrated by a solid line in FIG. 20 is the
.epsilon.-IVC valid range when the engine speed is 2,000 rpm, as
described above. If the engine speed changes, the .epsilon.-IVC
valid range also changes. When the engine speed increases, the
.epsilon.-IVC valid range is also parallelly translated to the
retard side, similarly in FIG. 20. Therefore, if the geometric
compression ratio is set as 10.ltoreq..epsilon.<20, when the
engine speed is 3,000 rpm (see a broken line), the ECU 10
determines the close timing IVC (deg.aBDC) so that the following
expression is satisfied.
0.9949.epsilon..sup.2-41.736.epsilon.+403.16.ltoreq.IVC.ltoreq.-0.9949.ep-
silon..sup.2+41.736.epsilon.-359.16 (11.sup.-1)
If the geometric compression ratio .epsilon. is set as
20.ltoreq..epsilon.<25, when the engine speed is 3,000 rpm, the
ECU 10 determines the close timing IVC (deg.aBDC) so that the
following expression is satisfied. -33.ltoreq.IVC.ltoreq.77
(12.sup.-1)
Further, if the geometric compression ratio .epsilon. is set as
25.ltoreq..epsilon..ltoreq.30, when the engine speed is 3,000 rpm,
the ECU 10 determines the close timing IVC (deg.aBDC) so that the
following expression is satisfied.
-4.7481.epsilon..sup.2+266.75.epsilon.-3669.2.ltoreq.IVC.ltoreq.77
(13.sup.-1) or
-33.ltoreq.IVC.ltoreq.4.7481.epsilon..sup.2-266.75.epsilon.+3713.2
(14.sup.-1)
If the geometric compression ratio .epsilon. is set as
10.ltoreq..epsilon..ltoreq.20, when the engine speed is 4,000 rpm
(see one-dot chain line), the ECU 10 determines the close timing
IVC (deg.aBDC) so that the following expression is satisfied.
0.9949.epsilon..sup.2-41.736.epsilon.+409.16.ltoreq.IVC.ltoreq.-0.9949.ep-
silon..sup.2+41.736.epsilon.-353.16 (11.sup.-2)
If the geometric compression ratio .epsilon. is set as
20.ltoreq..epsilon..ltoreq.25, when the engine speed is 4,000 rpm,
the ECU 10 determines the close timing IVC (deg.aBDC) so that the
following expression is satisfied. -27.ltoreq.IVC.ltoreq.83
(12.sup.-2)
Further, if the geometric compression ratio .epsilon. is set as
25.ltoreq..epsilon..ltoreq.30, when the engine speed is 4,000 rpm,
the ECU 10 determines the close timing IVC (deg.aBDC) so that the
following expression is satisfied.
-4.7481.epsilon..sup.2+266.75.epsilon.-3663.2.ltoreq.IVC.ltoreq.83
(13.sup.-2) or
-27.ltoreq.IVC.ltoreq.4.7481.epsilon..sup.2-266.75.epsilon.+3719.2
(14.sup.-2)
If the correction term C according to the engine speed NE (rpm) of
the engine 1 is used similar to the above, the relational
expression of .epsilon. and IVC in Layer 3 can be expressed as
follows. If the geometric compression ratio .epsilon. is
10.ltoreq..epsilon.<20,
0.9949.epsilon..sup.2-41.736.epsilon.+401.16+C.ltoreq.IVC.ltoreq.-0.9949.-
epsilon..sup.2+41.736.epsilon.-361.16+C (11.sup.-3) If the
geometric compression ratio .epsilon. is
20.ltoreq..epsilon..ltoreq.25, -35+C.ltoreq.IVC.ltoreq.75+C
(12.sup.-3) If the geometric compression ratio .epsilon. is
25.ltoreq..epsilon..ltoreq.30,
-4.7481.epsilon..sup.2+266.75.epsilon.-3671.2+C.ltoreq.IVC.ltoreq.75+C
(13.sup.-3) or
-35+C.ltoreq.IVC.ltoreq.4.7481.epsilon..sup.2-266.75.epsilon.+3711.2+C
(14.sup.-3)
The ECU 10 determines the close timing IVC based on the
.epsilon.-IVC valid range determined for every engine speed of the
engine 1. As a result, the ECU 10 can set the valve timing of the
intake valve 21 in Layer 3 as illustrated in FIG. 12.
Moreover, a lower figure 2002 of FIG. 20 is a relation between
.epsilon. and IVC when the fuel is the low octane fuel.
Upon determining the close timing IVC in the engine 1 of the low
octane fuel, if the geometric compression ratio .epsilon. is set as
10.ltoreq..epsilon.<18.7, when the engine speed is 2,000 rpm,
the ECU 10 determines the close timing IVC (deg.aBDC) so that the
following expression is satisfied.
0.9949.epsilon..sup.2-39.149.epsilon.+348.59.ltoreq.IVC.ltoreq.-0.9949.ep-
silon..sup.2+39.149.epsilon.-308.59 (15)
Moreover, if the geometric compression ratio .epsilon. is set as
18.7.ltoreq..epsilon.<23.7 in the engine 1 of the low octane
fuel, when the engine speed is 2,000 rpm, the ECU 10 determines the
close timing IVC (deg.aBDC) so that the following expression is
satisfied. -35.ltoreq.IVC.ltoreq.75 (16)
Further, if the geometric compression ratio .epsilon. is set as
23.7.ltoreq..epsilon..ltoreq.30 in the engine 1 of the low octane
fuel, when the engine speed is 2,000 rpm, the ECU 10 determines the
close timing IVC (deg.aBDC) so that the following expression is
satisfied.
-3.1298.epsilon..sup.2+172.48.epsilon.-2300.ltoreq.IVC.ltoreq.75
(17) or
-35.ltoreq.IVC.ltoreq.3.1298.epsilon..sup.2-172.48.epsilon.+2340
(18)
The cross-hatched range in the lower figure 2002 of FIG. 20 is an
overlapping range of the .epsilon.-IVC valid range of the high
octane fuel and the .epsilon.-IVC valid range of the low octane
fuel. Similar to the above, the control logic which suits both the
engine 1 using the high octane fuel and the engine 1 using the low
octane fuel is set by determining IVC within the overlapping range
of the two occurring ranges.
Note that although illustration is omitted, if the engine speed
increases, the .epsilon.-IVC valid range is parallelly translated
to the retard side also in the engine 1 of the low octane fuel. If
the geometric compression ratio .epsilon. is set as
10.ltoreq..epsilon.<18.7 in the engine 1 of the low octane fuel,
when the engine speed is 3,000 rpm, the ECU 10 determines the close
timing IVC (deg.aBDC) so that the following expression is
satisfied.
0.9949.epsilon..sup.2-39.149.epsilon.+350.59.ltoreq.IVC.ltoreq.-0.9949.ep-
silon..sup.2+39.149.epsilon.-306.59 (15.sup.-1)
Moreover, if the geometric compression ratio .epsilon. is set as
18.7.ltoreq..epsilon.<23.7 in the engine 1 of the low octane
fuel, when the engine speed is 3,000 rpm, the ECU 10 determines the
close timing IVC (deg.aBDC) so that the following expression is
satisfied. -33.ltoreq.IVC.ltoreq.77 (16.sup.-1)
Further, if the geometric compression ratio .epsilon. is set as
23.7.ltoreq..epsilon..ltoreq.30 in the engine 1 of the low octane
fuel, when the engine speed is 3,000 rpm, the ECU 10 determines the
close timing IVC (deg.aBDC) so that the following expression is
satisfied.
-3.1298.epsilon..sup.2+172.48.epsilon.-2298.ltoreq.IVC.ltoreq.77
(17.sup.-1) or
-33.ltoreq.IVC.ltoreq.3.1298.epsilon..sup.2-172.48.epsilon..sup.2+2342
(18.sup.-1)
If the geometric compression ratio .epsilon. is set as
10.ltoreq..epsilon.<18.7 in the engine 1 of the low octane fuel,
when the engine speed is 4,000 rpm, the ECU 10 determines the close
timing IVC (deg.aBDC) so that the following expression is
satisfied.
0.9949.epsilon..sup.2-39.149.epsilon.+356.59.ltoreq.IVC.ltoreq.-0.9949.ep-
silon..sup.2+39.149.epsilon.-300.59 (15.sup.-2)
Moreover, if the geometric compression ratio .epsilon. is set as
18.7.ltoreq..epsilon.<23.7 in the engine 1 of the low octane
fuel, when the engine speed is 4,000 rpm, the ECU 10 determines the
close timing IVC (deg.aBDC) so that the following expression is
satisfied. -27.ltoreq.IVC.ltoreq.83 (16.sup.-2)
Further, if the geometric compression ratio .epsilon. is set as
23.7.ltoreq..epsilon..ltoreq.30 in the engine 1 of the low octane
fuel, when the engine speed is 4,000 rpm, the ECU 10 determines the
close timing IVC (deg.aBDC) so that the following expression is
satisfied.
-3.1298.epsilon..sup.2+172.48.epsilon.-2292.ltoreq.IVC.ltoreq.83
(17.sup.-2) or
-27.ltoreq.IVC.ltoreq.3.1298.epsilon..sup.2-172.48.epsilon.+2348
(18.sup.-2)
If the correction term C according to the engine speed NE (rpm) of
the engine 1 is used similar to the above, the relational
expression of .epsilon. and IVC in Layer 3, in the engine 1 of the
low octane fuel can be expressed as follows. If the geometric
compression ratio .epsilon. is 10.ltoreq..epsilon.<18.7,
0.9949.epsilon..sup.2-39.149.epsilon.+348.59+C.ltoreq.IVC.ltoreq.-0.9949.-
epsilon..sup.2+39.149.epsilon.-308.59+C (15.sup.-3) If the
geometric compression ratio .epsilon. is
18.7.ltoreq..epsilon.<23.7, -35+C.ltoreq.IVC.ltoreq.75+C
(16.sup.-3) If the geometric compression ratio .epsilon. is
23.7.ltoreq..epsilon..epsilon.30,
-3.1298.epsilon..sup.2+172.48.epsilon.-2300+C.ltoreq.IVC.ltoreq.75+C
(17.sup.-3) or
-35+C.ltoreq.IVC.ltoreq.3.1298.epsilon..sup.2-172.48.epsilon.+2340+C
(18.sup.-3) (1-4) Relation Between Geometric Compression Ratio and
Close Timing of Intake Valve in Layers 2 and 3
FIG. 21 illustrates a relation between the geometric compression
ratio .epsilon. and the close timing IVC of the intake valve 21
where the SPCCI combustion is possible in both Layer 2 and Layer 3.
This relational expression is obtained from the .epsilon.-IVC valid
range of FIG. 17 and the .epsilon.-IVC valid range of FIG. 20.
When the ECU 10 selects Layer 3 according to the temperature, etc.
of the engine 1, the low-load operating range of the engine 1 is
switched from Layer 2 to Layer 3. If the close timing IVC of the
intake valve 21 is set so that the SPCCI combustion is possible in
both Layer 2 and Layer 3, it is possible to continuously perform
the SPCCI combustion even when the map of the engine 1 is switched
from Layer 2 to Layer 3.
An upper figure 2101 of FIG. 21 is a relation between .epsilon. and
IVC when the fuel is the high octane fuel. A lower figure 2102 is a
relation between .epsilon. and IVC when the fuel is the low octane
fuel.
If the geometric compression ratio .epsilon. is set as
10.ltoreq..epsilon.<17 in the engine 1 of the high octane fuel,
when the engine speed is 2,000 rpm, the ECU 10 determines the close
timing IVC (deg.aBDC) so that the following expression is
satisfied.
0.9949.epsilon..sup.2-41.736.epsilon.+401.16.ltoreq.IVC.ltoreq.-0.9949.ep-
silon..sup.2+41.736.epsilon.-361.16 (19)
If the geometric compression ratio .epsilon. is set as
17.ltoreq..epsilon..ltoreq.30 in the engine 1 of the high octane
fuel, when the engine speed is 2,000 rpm, the ECU 10 determines the
close timing IVC (deg.aBDC) so that the following expression is
satisfied.
-0.4288.epsilon..sup.2+31.518.epsilon.-379.88.ltoreq.IVC.ltoreq.-0.9949.e-
psilon..sup.2+41.736.epsilon.-361.16 (20) or
0.9949.epsilon..sup.2-41.736.epsilon.+401.16.ltoreq.IVC.ltoreq.1.9163.eps-
ilon..sup.2-89.935.epsilon.+974.94 (21)
By setting the close timing IVC of the intake valve 21 based on the
relational expressions (19) to (21), the SPCCI combustion of the
mixture gas with the A/F being leaner than the stoichiometric air
fuel ratio can be carried out, and the SPCCI combustion of the
mixture gas with the A/F being the stoichiometric air fuel ratio or
richer than the stoichiometric air fuel ratio, and the G/F being
leaner than the stoichiometric air fuel ratio can be carried
out.
Note that the close timing IVC is set for each operating state
which is determined by the load and the engine speed in Layer 2 and
Layer 3.
As illustrated by a broken line, if the geometric compression ratio
.epsilon. is set as 10.ltoreq..epsilon.<17 in the engine 1 of
the high octane fuel, when the engine speed is 3,000 rpm, the ECU
10 determines the close timing IVC (deg.aBDC) so that the following
expression is satisfied.
0.9949.epsilon..sup.2-41.736.epsilon.+403.16.ltoreq.IVC.ltoreq.-0.9949.ep-
silon..sup.2+41.736.epsilon.-359.16 (19.sup.-1)
If the geometric compression ratio .epsilon. is set as
17.ltoreq..epsilon..ltoreq.30 in the engine 1 of the high octane
fuel, when the engine speed is 3,000 rpm, the ECU 10 determines the
close timing IVC (deg.aBDC) so that the following expression is
satisfied.
-0.4288.epsilon..sup.2+31.518.epsilon.-377.88.ltoreq.IVC.ltoreq.-0.9949.e-
psilon..sup.2+41.736.epsilon.-359.16 (20.sup.-1) or
0.9949.epsilon..sup.2-41.736.epsilon.+403.16.ltoreq.IVC.ltoreq.1.9163.eps-
ilon..sup.2-89.935.epsilon.+976.94 (21.sup.-1)
Similarly, as illustrated by a one-dot chain line, if the geometric
compression ratio .epsilon. is set as 10.ltoreq..epsilon.<17 in
the engine 1 of the high octane fuel, when the engine speed is
4,000 rpm, the ECU 10 determines the close timing IVC (deg.aBDC) so
that the following expression is satisfied.
0.9949.epsilon..sup.2-41.736.epsilon.+409.16.ltoreq.IVC.ltoreq.-0.9949.ep-
silon..sup.2+41.736.epsilon.-353.16 (19.sup.-2)
If the geometric compression ratio .epsilon. is set as
17.ltoreq..epsilon..ltoreq.30 in the engine 1 of the high octane
fuel, when the engine speed is 4,000 rpm, the ECU 10 determines the
close timing IVC (deg.aBDC) so that the following expression is
satisfied.
-0.4288.epsilon..sup.2+31.518.epsilon.-371.88.ltoreq.IVC.ltoreq.-0.9949.e-
psilon..sup.2+41.736.epsilon.-353.16 (20.sup.-2) or
0.9949.epsilon..sup.2-41.736.epsilon.+409.16.ltoreq.IVC.ltoreq.1.9163.eps-
ilon..sup.2-89.935.epsilon.+982.94 (21)
If the correction term C according to the engine speed NE (rpm) of
the engine 1 is used similar to the above, the relational
expression of .epsilon. and IVC in Layer 2 and Layer 3 can be
expressed as follows. If the geometric compression ratio .epsilon.
is 10.ltoreq..epsilon.<17,
0.9949.epsilon..sup.2-41.736.epsilon.+401.16+C.ltoreq.IVC.ltoreq.-0.9949.-
epsilon..sup.2+41.736.epsilon.-361.16+C (19.sup.-3) If the
geometric compression ratio .epsilon. is
17.ltoreq..epsilon..ltoreq.30,
-0.4288.epsilon..sup.2+31.518.epsilon.-379.88+C.ltoreq.IVC.ltoreq.-0.9949-
.epsilon..sup.2+41.736.epsilon.-361.16+C (20.sup.-3) or
0.9949.epsilon..sup.2-41.736.epsilon.+401.16+C.ltoreq.IVC.ltoreq.1.9163.e-
psilon..sup.2-89.935.epsilon.+974.94+C (21.sup.-3)
Here, if the geometric compression ratio .epsilon. determined to be
lower than 17, the ECU 10 can determine IVC based on the relational
expression (19.sup.-3). Since the selection range of IVC is wide, a
degree of freedom in the design becomes high.
Moreover, as illustrated in the lower figure 2102 of FIG. 21, if
the geometric compression ratio .epsilon. is set as
10.ltoreq..epsilon.<15.7 in the engine 1 of the low octane fuel,
when the engine speed is 2,000 rpm, the ECU 10 determines the close
timing IVC (deg.aBDC) so that the following expression is
satisfied.
0.9949.epsilon..sup.2-39.149.epsilon.+348.59.ltoreq.IVC.ltoreq.-0.9949.ep-
silon..sup.2+39.149.epsilon.-308.59 (22)
Moreover, if the geometric compression ratio .epsilon. is set as
15.7.ltoreq..epsilon..ltoreq.30 in the engine 1 of the low octane
fuel, when the engine speed is 2,000 rpm, the ECU 10 determines the
close timing IVC (deg.aBDC) so that the following expression is
satisfied.
-0.5603.epsilon..sup.2+34.859.epsilon.-377.22.ltoreq.IVC.ltoreq.-0.9949.e-
psilon..sup.2+39.149.epsilon.-308.59 (23) or
0.9949.epsilon..sup.2-39.149.epsilon.+348.59.ltoreq.IVC.ltoreq.1.9211.eps-
ilon..sup.2-85.076.epsilon.+862.01 (24)
Although illustration is omitted, if the geometric compression
ratio .epsilon. is set as 10.ltoreq..epsilon.<15.7 in the engine
1 of the low octane fuel, when the engine speed is 3,000 rpm, the
ECU 10 determines the close timing IVC (deg.aBDC) so that the
following expression is satisfied.
0.9949.epsilon..sup.2-39.149.epsilon.+350.59.ltoreq.IVC.ltoreq.-0.9949.ep-
silon..sup.2+39.149.epsilon.-306.59 (22.sup.-1)
Moreover, if the geometric compression ratio .epsilon. is set as
15.7.ltoreq..epsilon..ltoreq.30 in the engine 1 of the low octane
fuel, when the engine speed is 3,000 rpm, the ECU 10 determines the
close timing IVC (deg.aBDC) so that the following expression is
satisfied.
-0.5603.epsilon..sup.2+34.859.epsilon.-375.22.ltoreq.IVC.ltoreq.-0.9949.e-
psilon..sup.2+39.149.epsilon.-306.59 (23.sup.-1) or
0.9949.epsilon..sup.2-39.149.epsilon.+350.59.ltoreq.IVC.ltoreq.1.9211.eps-
ilon..sup.2-85.076.epsilon.+864.01 (24.sup.-1)
Although illustration is similarly omitted, if the geometric
compression ratio .epsilon. is set as 10.ltoreq..epsilon.<15.7
in the engine 1 of the low octane fuel, when the engine speed is
4,000 rpm, the ECU 10 determines the close timing IVC (deg.aBDC) so
that the following expression is satisfied.
0.9949.epsilon..sup.2-39.149.epsilon.+356.59.ltoreq.IVC.ltoreq.-0.9949.ep-
silon..sup.2+39.149.epsilon.-300.59 (22.sup.-2)
Moreover, if the geometric compression ratio .epsilon. is set as
15.7.ltoreq..epsilon..ltoreq.30 in the engine 1 of the low octane
fuel, when the engine speed is 4,000 rpm, the ECU 10 determines the
close timing IVC (deg.aBDC) so that the following expression is
satisfied.
-0.5603.epsilon..sup.2+34.859.epsilon.-369.22.ltoreq.IVC.ltoreq.-0.9949.e-
psilon..sup.2+39.149.epsilon.-300.59 (23.sup.-2) or
0.9949.epsilon..sup.2-39.149.epsilon.+356.59.ltoreq.IVC.ltoreq.1.9211.eps-
ilon..sup.2-85.076.epsilon.+870.01 (24.sup.-2)
If the correction term C according to the engine speed NE (rpm) of
the engine 1 is used similar to the above, the relational
expression of .epsilon. and IVC in Layer 2 and Layer 3 in the
engine 1 of the low octane fuel can be expressed as follows. If the
geometric compression ratio .epsilon. is
10.ltoreq..epsilon.<15.7,
0.9949.epsilon..sup.2-39.149.epsilon.+348.59+C.ltoreq.IVC.ltoreq.-0.9949.-
epsilon..sup.2+39.149.epsilon.-308.59+C (22.sup.-3) If the
geometric compression ratio .epsilon. is
15.7.ltoreq..epsilon..ltoreq.30,
-0.5603.epsilon..sup.2+34.859.epsilon.-377.22+C.ltoreq.IVC.ltoreq.-0.9949-
.epsilon..sup.2+39.149.epsilon.-308.59+C (23.sup.-3) or
0.9949.epsilon..sup.2-39.149.epsilon.+348.59+C.ltoreq.IVC<1.9211.epsil-
on..sup.2-85.076.epsilon.+862.01+C (24.sup.-3)
Note that although illustration is omitted, the ECU 10 may
determine IVC within an overlapping range of the .epsilon.-IVC
valid range of the upper FIG. 2101 and the .epsilon.-IVC valid
range of the lower figure 2102 of FIG. 21. Similar to the above,
the ECU 10 can set the control logic which suits both the engine 1
using the high octane fuel and the engine 1 using the low octane
fuel by determining IVC within the overlapping range of the two
occurring ranges.
(1-5) Procedure of Method of Implementing Control Logic
Next, a procedure of the method of controlling the engine 1 for
performing the SPCCI combustion will be described with reference to
a flowchart illustrated in FIG. 22. The ECU 10 can execute each
step and stores information on the .epsilon.-IVC valid range
illustrated in FIGS. 17, 20, and 21.
At Step S221 after the procedure starts, the ECU 10 first sets the
geometric compression ratio .epsilon.. The set value of the
geometric compression ratio .epsilon. may be inputted from an
exterior device into the ECU 10.
At the following Step S222, the ECU 10 sets the valve opening angle
of the intake valve 21, and the valve opening angle of the exhaust
valve 22. This corresponds to determining the cam shapes of the
intake valve 21 and the exhaust valve 22. The set values of the
valve opening angles of the intake valve 21 and the exhaust valve
22 may be inputted from an exterior device into the ECU 10. Thus, a
hardware configuration of the engine 1 can be identified at Steps
S221 and S222.
At Step S223, the ECU 10 sets the operating state comprised of the
load and the engine speed, and at the following Step S224, the ECU
10 selects IVC based on the stored .epsilon.-IVC valid range (FIGS.
17, 20, and 21).
Then, at Step S225, the computer determines whether the SPCCI
combustion can be achieved based on IVC set at Step S224. If the
determination at Step S225 is YES, this procedure shifts to Step
S226, and the ECU 10 controls the engine 1 according to the control
logic corresponding to the engine operating state so that the SPCCI
combustion is performed in the operating state set at Step S223. On
the other hand, if the determination at Step S225 is NO, this
procedure shifts to Step S227, and the ECU 10 controls the engine 1
according to the control logic corresponding to the engine
operating state so that the SI combustion is performed in the
operating state set at Step S223. Note that at Step S227, the ECU
10 may again set the close timing IVC of the intake valve 21 in
consideration of performing the SI combustion.
As described above, the control device of the compression-ignition
engine disclosed herein can set the close timing IVC of the intake
valve 21 within the range where the relation between the engine
geometric compression ratio .epsilon. and the close timing IVC of
the intake valve 21 is satisfied.
(2) Method of Controlling Supercharging Pressure
As described above, the close timing IVC of the intake valve 21 can
be set based on the relation between the geometric compression
ratio .epsilon. and the close timing IVC of the intake valve
21.
As a result of further repeated and diligent examinations, the
present inventors further found that, in order to put in practical
use the engine 1 for performing the SPCCI combustion with a lean
G/F in Layer 2, another relation was required between a
supercharging pressure P of the engine 1 and the close timing IVC
of the intake valve 21. The present inventors further found a
control method based on such another relation, after the
implementing method was utilized.
(2-1) Relation Between Supercharging Pressure and Close Timing of
Intake Valve in Layer 2
As described above, the engine 1 operates, while making the A/F of
the mixture gas the stoichiometric air fuel ratio or a
substantially stoichiometric air fuel ratio in Layer 2, and making
the G/F leaner than the stoichiometric air fuel ratio. Here, the
G/F can be controlled through the amount of gas supplied into the
combustion chamber 17, and a control factor for controlling the
amount of gas includes the supercharging pressure P of the
supercharger 44, in addition to an overlap period O/V.
For example, since the fuel supply amount is large when operating
the engine 1 in the high-load operating state including the maximum
load operating state, a larger amount of gas is required to be
supplied into the combustion chamber 17. Therefore, in order to
maintain the G/F lean in Layer 2, it may be necessary to set the
supercharging pressure P (especially, a target supercharging
pressure of the supercharger 44) at a given pressure or higher.
That is, in Layer 2, a lower limit of the supercharging pressure P
can be determined. Moreover, an actual amount of gas sent into the
combustion chamber 17 from the intake passage 40 among the gas
supercharged by the supercharger 44 depends also on the close
timing IVC of the intake valve 21.
Therefore, as a condition for maintaining the G/F at lean in Layer
2, a relation between the lower limit of the supercharging pressure
P and the close timing IVC can be determined.
Therefore, the present inventors estimated the G/F by using the
model of the engine 1 also in consideration of an inertia effect of
air, while changing the values of the two parameters of the
supercharging pressure P and the close timing IVC. Then, the
present inventors acquired a combination of the lower limit of the
supercharging pressure P and the close timing IVC so that the G/F
becomes the given value or higher.
FIG. 23 illustrates an approximation (A) calculated from the
combination of the lower limit of P and IVC. The horizontal axis of
FIG. 23 is the close timing IVC (deg. aBDC) of the intake valve 21,
and the vertical axis is the supercharging pressure P (kPa) of the
engine 1. In FIG. 23, "E.sup.-n" (n is an integer) indicates
"10.sup.-n." For example, "8E-11IVC.sup.6" means
"8.times.10.sup.-11.times.IVC.sup.6."
An upper figure of FIG. 23 corresponds to a case where the engine
speed of the engine 1 is 2000 rpm or higher. The approximation (A)
corresponds to a solid line in FIG. 23.
P=8.0.times.10.sup.-11IVC.sup.6-1.0.times.10.sup.-8IVC.sup.5+3.0.times.10-
.sup.-7IVC.sup.4-4.0.times.10.sup.-6IVC.sup.3+0.0068IVC.sup.2-0.3209IVC+11-
6.63 Approximation (A)
In FIG. 23, the combination of P and IVC above the approximation
(A) can achieve the SPCCI combustion, while maintaining the G/F at
lean, because the G/F becomes the given value or higher.
As seen in FIG. 23, the relation between P and IVC is substantially
laterally symmetrical with respect to IVC=20 deg.aBDC. IVC=20
deg.aBDC corresponds to a close timing (best IVC) at which the
amount of gas introduced into the combustion chamber 17 becomes the
maximum.
That is, at IVC=20 deg.aBDC, a large amount of gas can be supplied
into the combustion chamber 17, even if the supercharging pressure
P is set relatively low. Therefore, the supercharging pressure P is
the minimum at IVC=20 deg.aBDC. Moreover, in order to maintain the
G/F at a given value or higher, it is required to set the
supercharging pressure P relatively high as the G/F deviates from
this minimum. Therefore, as IVC deviates from 20 deg.aBDC, it is
necessary to increase the supercharging pressure P.
Note that the close timing IVC is set for each operating state
which is defined by the load and the engine speed in Layer 2. The
example of FIG. 23 is a P-IVC valid range when the engine speed of
the engine 1 is 2000 rpm, as described above. If the engine speed
of the engine 1 changes, the P-IVC valid range also changes. The
ECU 10 determines the close timing IVC based on the P-IVC valid
range defined for every engine speed of the engine 1.
The supercharging pressure P described above may also be a target
supercharging pressure of the supercharger 44 in Layer 2.
Moreover, it is generally thought that, as for the supercharging
pressure P which can be set to the supercharger 44, an upper limit
(a so-called supercharger limit) due to its hardware configuration
exists. Therefore, in Layer 2, the upper limit can be provided to
the supercharging pressure P.
FIG. 23 also illustrates the approximation (B) indicative of the
upper limit of P. The approximation (B) corresponds to the
reference character P1 in FIG. 23, and in the illustrated example,
P=150. Approximation (B)
In FIG. 23, the combination of P and IVC below the approximation
(B) permits the supercharging pressure P to be set within a range
without exceeding the supercharger limit. Moreover, although the
supercharger limit indicated by the approximation (B) varies
according to the hardware configuration of the supercharger 44, the
limit may be set so as to at least exceed the minimum of the
supercharging pressure P in the approximation (A).
Moreover, other than the upper limit of P based on the
approximation (B), an upper limit (second upper limit) of P in
consideration of the fuel efficiency may be provided. Thus, the
loss of work caused by excessive supercharging pressure can be
reduced, and this becomes advantageous to improve the fuel
efficiency of the engine. Although illustration is omitted, this
second upper limit can be set to, for example, P=250.
As illustrated in the above equation, the second upper limit can be
set higher than the upper limit illustrated in the approximation
(B).
Moreover, in order to achieve the SPCCI combustion in Layer 2, the
relation between .epsilon. and IVC is required in the first place
so that the temperature of the combustion chamber 17 at .theta.CI
becomes the reference temperature Tth2 or higher. Such a
requirement is satisfied by, for example, the combination of
.epsilon. and IVC on the right side of the approximations (III) and
(IV) in FIG. 17.
That is, in order to achieve the SPCCI combustion in Layer 2, IVC
needs to fall within a given range. That is, an upper limit and a
lower limit can be set to IVC.
Thus, when estimating the G/F, the present inventors set the upper
limit and the lower limit of IVC according to the value of the
geometric compression ratio .epsilon..
In FIG. 23, an approximation (C) indicative of the upper limit of
IVC and an approximation (D) indicative of the lower limit of IVC
are also illustrated. The approximation (C) is determined based on
the approximation (III) in FIG. 17, and corresponds to a reference
character Ir in FIG. 23. On the other hand, the approximation (D)
is determined based on the approximation (IV) in FIG. 17, and
corresponds to a reference character Il in FIG. 23. In the
illustrated example, the approximations (C) and (D) are as follows.
IVC=90 Approximation (C) IVC=-50 Approximation (D)
In FIG. 23, the combination of P and IVC on the left side of the
approximation (C) and on the right side of the approximation (D)
can stably achieve SPCCI combustion.
Note that the approximations (C) and (D) in the upper figure of
FIG. 23 correspond to a case where the geometric compression ratio
.epsilon. is set to around 16 as can be seen in FIG. 17. If
.epsilon. is set smaller than this, as illustrated in the lower
figure of FIG. 23, the upper limit and the lower limit of IVC both
approach IVC=20 deg.aBDC. Moreover, the approximations (C) and (D)
vary according to the octane number of the fuel.
Note that the approximation (A) in the lower figure of FIG. 23 is
made the same as that of the upper figure for the convenience of
illustration, but the details of the approximation (A) may vary in
fact according to the geometric compression ratio .epsilon..
The combination of P and IVC within a range surrounded by the
approximations (A), (B), (C), and (D) in FIG. 23 is the combination
which can achieve the SPCCI combustion, while maintaining the G/F
in Layer 2 at the given value or higher. In other words, the
combination of P and IVC outside this range cannot achieve the
G/F-lean SPCCI combustion in Layer 2.
The ECU 10 must determine IVC within the P-IVC valid range which is
hatched in FIG. 23, when operating the engine 1 in Layer 2 and
determining the close timing IVC of the intake valve 21.
Specifically, in the configuration example illustrated in the upper
figure of FIG. 23, the ECU 10 determines the close timing IVC
(deg.aBDC) so that the supercharging pressure P (kPa) satisfies the
following expression.
P.gtoreq.8.0.times.10.sup.-11IVC.sup.6-1.0.times.10.sup.-8IVC.sup.5+3.0.t-
imes.10.sup.-7IVC.sup.4-4.0.times.10.sup.-6IVC.sup.3+0.0068IVC.sup.2-0.320-
9IVC+116.63 (25)
Moreover, the ECU 10 determines the close timing IVC (deg.aBDC) so
that the supercharging pressure P (kPa) satisfies the following
expression. P.ltoreq.150 (26)
Moreover, the ECU 10 determines the close timing IVC (deg.aBDC) so
as to satisfy the following expression. -50.ltoreq.IVC.ltoreq.90
(27) (2-2) Procedure of Control Method
Next, a procedure of the method of controlling the engine 1 for
performing the SPCCI combustion will be described with reference to
a flowchart illustrated in FIG. 24. The ECU 10 can execute each of
the following steps. The ECU 10 stores information on the P-IVC
valid range illustrated in FIG. 23.
At Step S241 after the procedure starts, the ECU 10 first sets the
geometric compression ratio .epsilon.. A preset value of the
geometric compression ratio .epsilon. may be inputted into the ECU
10 from an exterior device.
At the subsequent Step S242, the ECU 10 sets a valve opening angle
of the intake valve 21, and a valve opening angle of the exhaust
valve 22. This corresponds to determining cam shapes of the intake
valve 21 and the exhaust valve 22. The preset values of the valve
opening angles of the intake valve 21 and the exhaust valve 22 may
be inputted into the ECU 10 from an exterior device. A hardware
configuration of the engine 1 can be specified at Steps S241 and
S242.
At Step S243, the ECU 10 sets the operating state comprised of the
load and the engine speed of the engine 1, and at the subsequent
Step S244, the ECU 10 sets a target supercharging pressure as the
supercharging pressure P based on the operating state set at Step
S243.
At the subsequent Step S245, the ECU 10 determines IVC based on the
prestored P-IVC valid range (FIG. 23).
Then, at Step S246, the ECU 10 determines whether SPCCI combustion
can be achieved based on IVC set at Step S245. If the determination
of Step S246 is YES, the procedure shifts to Step S247, where the
ECU 10 controls the engine 1 according to the control logic
corresponding to this operating state so that the SPCCI combustion
is performed in the operating state set at Step S243. On the other
hand, if the determination of Step S246 is NO, the procedure shifts
to Step S248, where the ECU 10 controls the engine 1 according to
the control logic corresponding to this operating state so that the
SI combustion is performed in the operating state set at Step S243.
Note that at Step S245, the ECU 10 may again set the close timing
IVC of the intake valve 21 in consideration of performing the SI
combustion.
As described above, in the control device of the
compression-ignition engine disclosed herein, the relation between
the supercharging pressure P of the engine 1 and the close timing
IVC of the intake valve 21 is defined beforehand, and the ECU 10
can set the close timing IVC of the intake valve 21 within a range
where this relation is satisfied.
(Other Embodiments)
Note that the application of the technology disclosed herein is not
limited to the engine 1 having the configuration described above.
The engine 1 may adopt various configurations.
For example, the engine 1 may be provided with a turbocharger,
instead of the mechanical supercharger 44.
It should be understood that the embodiments herein are
illustrative and not restrictive, since the scope of the invention
is defined by the appended claims rather than by the description
preceding them, and all changes that fall within metes and bounds
of the claims, or equivalence of such metes and bounds thereof, are
therefore intended to be embraced by the claims.
DESCRIPTION OF REFERENCE CHARACTERS
1 Engine
10 ECU (Control Part)
17 Combustion Chamber
23 Intake-side Electric S-VT (Variable Valve Operating
Mechanism)
25 Ignition Plug (Ignition Part)
44 Supercharger
55 EGR System
6 Injector (Fuel Injection Part)
SW1 Airflow Sensor (Measurement Part)
SW2 First Intake-air Temperature Sensor (Measurement Part)
SW3 First Pressure Sensor (Measurement Part)
SW4 Second Intake-air Temperature Sensor (Measurement Part)
SW5 Second Pressure Sensor (Measurement Part)
SW6 Pressure Indicating Sensor (Measurement Part)
SW7 Exhaust Temperature Sensor (Measurement Part)
SW8 Linear O.sub.2 Sensor (Measurement Part)
SW9 Lambda O.sub.2 Sensor (Measurement Part)
SW10 Water Temperature Sensor (Measurement Part)
SW11 Crank Angle Sensor (Measurement Part)
SW12 Accelerator Opening Sensor (Measurement Part)
SW13 Intake Cam Angle Sensor (Measurement Part)
SW14 Exhaust Cam Angle Sensor (Measurement Part)
SW15 EGR Pressure Difference Sensor (Measurement Part)
SW16 Fuel Pressure Sensor (Measurement Part)
SW17 Third Intake-air Temperature Sensor (Measurement Part)
* * * * *